High density combined cycle power plant process

ABSTRACT

A process for increasing the specific output of a combined cycle power plant and providing flexibility in the power plant rating, both without a commensurate increase in the plant heat rate, is disclosed. The present invention demonstrates that the process of upgrading thermal efficiencies of combined cycles can often be accomplished through the strategic use of additional fuel and/or heat input. In particular, gas turbines that exhaust into HRSGs, can be supplementally fired to obtain much higher steam turbine outputs and greater overall plant ratings, but without a penalty on efficiency. This method by and large defines a high efficiency combined cycle power plant that is predominantly a Rankine (bottoming) cycle. Exemplary embodiments of the present invention include a load driven by a topping cycle engine, powered by a topping cycle fluid which exhausts into a heat recovery device.

RELATED APPLICATIONS

[0001] This is a divisional of application Ser. No. 09/783,693, filed,Feb. 14, 2001 which was a divisional of application Ser. No. 09/359,813,filed Jul. 23, 1999, which claimed the benefit of U.S. ProvisionalApplication Nos. 60/098,468, filed Aug. 31, 1998, and 60/125,576, filedMar. 23, 1999.

FIELD OF THE INVENTION

[0002] This invention relates generally to combined cycle power plantsthat may or may not incorporate cogeneration into their cycle. As willbe demonstrated by the following disclosure, the increasing need formore energy efficient and environmentally friendly methods of generatingpower has prompted a widespread search for systems and methods toachieve these goals. However, current technologies have a generallymyopic view of the total economic impact imposed by a concentration onenergy efficiency and environmental issues alone.

[0003] The present invention proposes to break with tradition andinclude as part of the economic and environmental analysis the completeequipment complement required to implement a desired plant load (power)rating. By incorporating this analysis into a new system and method ofsupplemental firing and heat recovery, the present inventiondramatically cuts the overall economic and environmental cost ofinstalled power plants by reducing the equipment complement whilemaintaining or reducing plant emissions. The result of this improvementover the art is cheaper and cleaner electrical energy than would bepossible using conventional combined cycle plants that are currentlyknown in the art.

BACKGROUND OF THE INVENTION

[0004] Overview

[0005] Combined cycle power plants and cogeneration facilities utilizegas turbines (GT(s)) as prime movers to generate power. These GT enginesoperate on the Brayton Cycle thermodynamic principle and typically havehigh exhaust flows and relatively high exhaust temperatures. Theseexhaust gases, when directed into a heat recovery boiler (typicallyreferred to as a heat recovery steam generator (HRSG)), produce steamthat can be used to generate more power and/or provide process steamrequirements. For additional power production the steam can be directedto a steam turbine (ST) that utilizes the steam to produce additionalpower. In this manner, the GT produces work via the Brayton Cycle, andthe ST produces power via the Rankine Cycle. Thus, the name “combinedcycle” is derived. In this arrangement, the GT Brayton Cycle is alsoreferred to as the “topping cycle” and the ST Rankine Cycle is referredto as the “bottoming cycle,” as the topping cycle produces the energyneeded for the bottoming cycle to operate. Thus, the functionality ofthese cycles is linked in the prior art.

[0006] Rankine Cycle

[0007] Steam has been used for power applications for more than acentury. Early applications utilized a pump to bring the water up to thedesired pressure, a boiler to heat the water until it turned to steam,and a steam engine, typically a piston type engine, to produce shafthorsepower. These power plants were used in factories, on locomotives,onboard steamships, and other power applications.

[0008] As technology progressed, the trend for the use of steam enginesdiminished and the use of steam turbines increased. One advantage of thesteam turbine was its overall cycle efficiency when used in conjunctionwith a condenser. This allowed the steam to expand significantly beyondnormal atmospheric pressure down to pressures that were only slightlyabove an absolute vacuum (0.5 to 2 pounds per square inch absolute(psia)). This allowed the steam to expand further than in an atmosphericexhaust configuration, extracting more energy from a given mass ofsteam, thus producing more power and increasing overall steam cycleefficiency. This overall steam cycle, from a thermodynamic perspective,is referred to as the Rankine Cycle.

[0009]FIG. 1 illustrates the thermodynamic operation of the RankineCycle. In FIG. 1, graph (100) illustrates the Rankine Cycle on aPressure versus Volume plot. From point (101) to point (102), water ispressurized at constant volume. From point (102) to point (103), thewater is boiled into steam at constant pressure. Point (103) to point(104) defines the process where the steam expands isentropically andproduces work. Then, from point (104) to point (101) the low-pressuresteam is condensed back to water and the cycle is complete.

[0010] Also in FIG. 1, graph (110) illustrates the Rankine Cycle on aTemperature versus Entropy plot. From point (111) to point (112), wateris pressurized. From point (112), the water is boiled into steam atconstant temperature until it is all steam, then it is superheated topoint (113). Point (113) to point (114) defines the process where thesteam expands isentropically and produces work. From point (114) topoint (111) the low-pressure steam is condensed back to water atconstant temperature to complete the cycle. See Eugene A. Avallone andTheodore Baumeister III, MARKS' STANDARD HANDBOOK FOR MECHANICALENGINEERS (NINTH EDITION) (ISBN 0-07-004127-X, 1987) in Section 4-20 formore discussion on the Rankine Cycle.

[0011] Power Plant Cycle

[0012] For a number of decades, the Rankine Cycle has been used toproduce most of the electricity in the United States, as well as in anumber of other countries. FIG. 2 illustrates a schematic of the basicRankine Cycle, with the four primary components being the Boiler FeedPump (BFP) (201), Boiler evaporator/superheater (BOIL) (203, 205), SteamTurbine (ST) (207), and the Condenser (COND) (209). Note that either oneor multiples of any component are possible in the arrangement, but forsimplicity, only one of each is shown in FIG. 2. The sub-criticalRankine Cycle (steam pressures less than 3206.2 psia) starts as water atthe inlet (211) of the BFP (201). The water is then pumped to a desireddischarge pressure by the BFP (201). This pressurized water (202) isthen sent to the evaporator (EVAP) (203) where heat is added to thepressurized water. Typically this is accomplished by burning a fuel inthe boiler, and the heat of combustion is then transferred to thepressurized water that is routed through tubes and other passages and/orvessels in the boiler. As sufficient heat is added to the pressurizedwater, it boils and turns into steam (204). This steam now exists in thetwo-phase region where both steam and water coexist at the same pressureand temperature, called the saturation pressure and saturationtemperature. For most applications designed in recent decades, thissteam (204) is then sent to a superheater section (SHT) (205) in theboiler where it is heated to a higher temperature than saturationtemperature. This steam (206) is now referred to as superheated steam.Superheated steam reduces (but does not eliminate) the risk of watercarryover into the steam turbine (207), which is of concern since watercarryover can cause extensive internal steam turbine damage. Of moreimportance, however, is the fact that superheated steam yields bettercycle efficiencies. This is of great importance to large central powerstations.

[0013] Once produced, the superheated steam (206) is sent to the steamturbine (207), typically via one or more pipes. The steam then begins toexpand in the steam turbine (ST) and produce shaft horsepower. Aftertraveling through the steam turbine down to a low exhaust pressure, thesteam exits the ST (208), and is sent to the condenser (209), where itis then condensed back into water. This device is typically a tubed heatexchanger, but can also be other types of heat exchangers such as aspray chamber, air-cooled condenser, or other heat exchange device usedfor a similar purpose. After rejecting heat from the low-pressure steamand condensing the steam back to water, the condenser collects the waterin an area commonly referred to as the hotwell (HW) (210), where it isthen typically pumped through the condensate line (211) and back to theBFP (201). Shaft horsepower produced in the ST is converted intoelectrical power in the generator (GEN) (212). This cycle of one unit ofwater from the point of beginning, through the system, and back to thepoint of origin defines the basic Rankine Cycle.

[0014] Current power plants using only steam as the motive fluidtypically use a boiler to produce the steam. This boiler may be fueledby a variety of fuels, including oil, natural gas, coal, biomass, aswell as others, such as nuclear fuel. The boilers may also use acombination of fuels as well. Depending upon capital costconsiderations, fuel costs, maintenance issues, and other factors, theowners and engineers will select the steam pressure and temperature atwhich the boiler will produce steam.

[0015] Due to the size and weight of large steam turbines, they requireextended periods for start-up. This is due to the thick metal casingsand large heavy rotors that are utilized in their construction.Therefore, these machines require long start-up periods to allow theseheavy components to warm up uniformly, and avoid interference betweenstationary and rotating parts that may occur due to differential thermalexpansion.

[0016] Although the heavy construction is a deterrent to rapid startup,it provides for robust construction and sustained performance levels.Even after four (4) years of nearly continuous service, the performancedecay for a large ST should be less than 2%. This performance decay,combined with the fact that the boiler feed pumps only consume about 2%of the ST output, mean that the performance levels for a ST sustain nearoptimum levels for extended periods of time, even with decay in theauxiliary loads (BFP). In other words, if the BFP efficiency decays from75% to 65%, the auxiliary load only increases from 2.00% to 2.31%. Thisis a small effect on the net output of the Rankine cycle plant, and isanother one of its major advantages.

[0017] Brayton Cycle

[0018] The Brayton Cycle varies quite differently from the RankineCycle, as a major part of the cycle involves the compression of theworking fluid, which is a compressible gas. This process consumes agreat deal of power, therefore, efficient compression of the workingfluid is essential to an efficient Brayton Cycle.

[0019] Common engines that utilize a Brayton Cycle are aircraftturboprops, jet engines, and gas turbines for stationary application.These engines work by ingesting air (the working fluid), compressing itto a higher pressure, typically 3 to 30 times that of the surroundingambient air, adding heat through direct combustion (although heataddition from an external source is also possible), and then expandingthe resulting high-pressure hot gases through a turbine section.Aircraft engines primarily produce thrust to propel an aircraft throughthe air. Therefore, some or perhaps none of their output is in the formof shaft horsepower (a turboprop gas turbine engine may drive thepropeller, but may also produce some thrust from the high velocityexhaust gases).

[0020] For stationary gas turbine applications, the purpose of theengine is to produce shaft horsepower. Approximately ⅔ of the energyproduced by the turbine section of the gas turbine is required to drivethe compressor section, with the remaining ⅓ available to drive a load.This drawback of GT systems may be used to advantage in the presentinvention as described later in this document.

[0021] Aircraft engines utilize the Brayton Cycle because these enginesoffer high thrust-to-weight ratios. This is needed to minimize theaircraft weight so it can fly. For stationary applications, gas turbinesare used to provide electrical power at peak loads. This is anotheradvantage the Brayton Cycle engines have over Rankine Cycle engines:rapid start and stop times (relatively speaking). Since steam turbinesare large heavy engines, it is necessary to start them slowly, and allowthe heat to slowly soak into the thick casings so as to avoid thermaldistortion and potential rubs between the stationary components androtating components of the engine. A large power plant steam turbine mayrequire a 24-hour warm-up sequence from cold start to reach full load.However, due to the lower operating pressures and lighter weights, gasturbines can be started and brought to fall load within a matter ofminutes of start-up.

[0022] Therefore, many utilities in the United States and othercountries use gas turbines to provide electrical power during peakdemand. These turbines are not very efficient in simple cycle (25% to30% LHV), but meet the electrical demand requirements for a few hourseach day.

[0023] Steam Turbine Design

[0024] When designing a steam turbine for a power plant application(constant speed), the steam turbine design engineer first examines theoutput rating desired by the customer. This is because the steam turbinewill be custom designed and manufactured for the customer to hisspecification. The steam turbine will not be totally designed from aclean sheet of paper as may be inferred by “custom”, but will utilizecomponents from a “family” of hardware and have a unique steam path forthe application. After turbine rating, the ST design engineer will lookat the plant steam conditions, and based upon these parameters determinean inlet flow to the turbine high-pressure (HP) section. Utilizing thisinformation, the ST design engineer can select the optimum HP casing forthe application. In a similar fashion, he can also select the optimumintermediate pressure (IP) and low-pressure (LP) casings as well.

[0025] Knowing which casings to use, the engineer then selects theappropriate blading (both stationary and rotating) for the application.This blading size is determined primarily by the volume flow (as opposedto mass flow) of steam through the turbine. With casings and bladingdetermined, the engineer completes the ST design by selecting valves,controls, instrumentation, and other accessories required for operationof the ST. The final design is a high efficiency ST optimized for thecustomer's steam conditions and desired rating.

[0026] An interesting note concerning this design philosophy is that twoSTs with the same steam conditions but with large differences in rating(for example, 200 MW versus 400 MW) may actually appear almost identicalwhen viewed from the outside. This is because the optimum casingsselected were designed to cover the flow range of both units. However,due to the large volume flow differences, the large unit would haveblades that are approximately twice the size (height) internally. It isinteresting to note, however, that both these units might have nearlythe same HP and IP casings. This means that the larger ST, even with adramatic increase in rating, may be only incrementally more expensive tomanufacture than the ST with the lower rating. This fact may be used toadvantage in the present invention as described later in this document.

[0027] Gas Turbine Design

[0028] Unlike the steam turbine, the gas turbine is not a customdesigned machine for each customer. Although accessories such as thestarting means, lube oil cooler type, and control options may bespecified by the customer for a particular application, the core engineis essentially standard. Much of this is due to the fact that the gasturbine is actually a packaged power plant, which needs essentially onlyfuel to produce power. In contrast, the steam turbine is merely acomponent of a power plant, and requires a boiler, BFP, and condenser tobecome a complete power plant. Therefore, the gas turbine compressorsection, combustion system, and turbine section must all be designed towork together. Since the design of the GT is a highly intensiveengineering task, GT designs are generally completed and extensivelytested, after which they are mass produced without variation to the coreengine design. This eliminates the customer's ability to specify poweroutput for either a facility with gas turbines only or a combined cyclefacility in the prior art. When building a combined cycle plant, thecustomer simply must choose from a selection of standard offerings by amanufacturer that best meets his needs for power output, efficiency, andcost.

[0029] Steam Turbine/Gas Turbine Efficiency and Rating Comparison

[0030] The largest and most efficient GT available today for 60-cyclepower production is rated at approximately 250 MW with an efficiency of40.0% LHV (Lower Heating Value). An example of this GT is theWestinghouse model 501G. This is in contrast to STs that can be rated upto as high as 1500 MW and have overall cycle efficiencies in excess of45% LHV. Therefore, comparing a Rankine Cycle power plant to a Braytoncycle power plant, where each employs the largest and most efficientturbine available, the single ST Rankine cycle is approximately six (6)times larger in rating and 12.5% more efficient than the Brayton Cyclewith its best GT. This fact may be used to advantage in the presentinvention as described later in this document.

[0031] Cogeneration/Combined Cycle

[0032] One characteristic of the gas turbine is that it expels highvolumes of exhaust gases at high temperature. With the advent of theArab oil embargo of 1973 and higher energy prices, more focus was put onfinding ways to utilize the energy contained in these high temperatureexhaust gases.

[0033] Significantly higher energy prices in the early 1970s signaledthe start of a wave of small power plants built using the principles ofcogeneration. Cogeneration is defined as the simultaneous production ofmechanical or electrical energy in conjunction with thermal energy. Inother words, the utilization of an engine (gas turbine or otherwise) toproduce power, while at the same time using waste heat from the enginefor another process, thus displacing fuel that would otherwise be usedfor said process. This was a very efficient method from a fuelutilization perspective and was encouraged by the United States PublicUtilities Regulation and Policies Act (PURPA) of 1978, which mandatedthat the local utilities must purchase power from qualifiedcogenerators, and buy it at a rate which included avoided cost for newpower plants.

[0034] At first cogeneration projects were small, typically less than 50MW. They consisted of small gas turbines with a HRSG to produce steam.In many instances, the steam pressures were relatively low (less than600 psig), as the steam was used for process requirements. Some projectsincluded a steam turbine, while others did not. As the industry matured,larger plants with higher steam pressures were designed to increasebottoming cycle efficiency. In addition, the major gas turbinemanufacturers designed and built larger and more efficient gas turbinesto meet the needs of the cogeneration marketplace. Soon, due to theirhigh efficiency, low emissions, and low capital cost (dollars per kW ofcapacity), cogeneration power plants gave way to combined cycle powerplants (plants that produced only power and provided no useful thermalenergy as was the case with cogeneration plants). Some cogenerationprojects are still being proposed and constructed, but they are nowtypically referred to as combined heat and power (CHP) projects.

[0035] Although there was this gradual shift from small cogenerationprojects to large combined cycle power plants, the arrangement andoverall system and method for producing power was for the most partunchanged. The gas turbine(s) was the primary engine, and a HRSG wasutilized to capture the heat in the GT exhaust gases. Optimized formaximum power production, the steam turbine(s) produced additional powerequal to approximately 50% of the power produced by the gas turbine(s).The HRSG was typically a two or three pressure level boiler to maximizeheat recovery and steam turbine was designed to accept steam from allpressure levels of the HRSG. A review of the manufacturers standardcombined cycle offerings will illustrate this trend. The 1997TURBOMACHINERY HANDBOOK, (USPS 871-500, ISSN 0149-4147), tabulatesstandard combined cycle power plants available from variousmanufacturer's including ABB, General Electric, and Westinghouse. Inmost every instance, the steam turbine's output is within the range of40% to 60% of the gas turbine(s) output. General Electric informativedocument GER-3567G, 1996, “GE Heavy-Duty Gas Turbine PerformanceCharacteristics,” by Frank J. Brooks provides the output for the gasturbines used in their combined cycle power plants.

[0036] In summary, the system and method utilized by the majormanufacturer's of combined cycle power plant turbomachinery evolved fromthe small cogeneration power facilities that were designed to produceboth power and thermal energy simultaneously. The sizes for combinedcycle power plants have grown from small cogeneration projects under 50MW to large structured plants producing in excess of 700 MW (as in theWestinghouse 2X1 501G combined cycle). These plants are primarily gasturbine power plants, with the steam turbine producing additional powerwhich is nominally 40% to 60% of the power produced by its associatedgas turbine(s). With the gas turbine as the prime engine, the ratings onthe standard combined cycle power plants are very rigid, as gas turbinesare production line items, versus steam turbines which are largelycustom designed and manufactured. A new system and method that offersmore flexibility, without compromising the benefits of combined cyclepower such as high efficiency, low emissions, and low capital cost,would be welcomed by the industry.

DESCRIPTION OF THE PRIOR ART

[0037] Efficiency Optimizations

[0038] Feedwater Heater

[0039] With Rankine Cycle plants producing billions of dollars ofelectricity annually, and consuming commensurate amounts of fuel eachyear, a great deal of design and analysis has been done to optimize theRankine Cycle by introducing small variations or revised configurations.FIG. 3 illustrates some of the common variations that are used to designa Rankine Cycle for optimum efficiency. Part (303) of FIG. 3schematically represents a feedwater heater (FWH). This device istypically a shell and tube heat exchanger, but could be a plate andframe heat exchanger, vortex mixing heat exchanger that mixes thefeedwater with small amounts of steam, or other heat exchange deviceused for a similar purpose. Analysis has proven that utilizingextraction steam from the steam turbine to preheat water before itenters the boiler increases the cycle efficiency.

[0040] The feedwater heater (303) uses steam that is extracted from thesteam turbine at an optimum point to preheat the water between thecondenser (319) outlet and the boiler inlet (306). A second feedwaterheater (305) is shown in this example. The number of feedwater heatersand their optimum steam conditions are dependent upon a number offactors including but not limited to steam turbine inlet pressure, steamturbine inlet temperature, reheat steam conditions, feedwater heatereffectiveness, and other factors. Typically, the number of feedwaterheaters, their design, and the inlet steam conditions for thesefeedwater heaters must be determined for each power plant due tovariations in each power plant's design and individual conditions.

[0041] Reheat

[0042] Another variation on the Rankine Cycle used to improve cycleefficiency is the use of reheat. This variation involves expanding steamin the steam turbine from design inlet conditions down to some specifiedreheat pressure. At this point, some energy has already been extractedfrom the steam to produce shaft horsepower. This lower energy contentsteam is then redirected to the boiler where it is reheated to a highertemperature. This higher energy content steam is then sent back to thesteam turbine to produce more power. More than one reheat can beutilized in the cycle. Again, for the given design conditions, inletpressures, inlet temperatures, and other conditions, the reheat isdesigned for the greatest benefit and increase in cycle efficiency.

[0043] Other Factors

[0044] Other factors that affect cycle efficiency include inlet steampressure, inlet steam temperature, and exhaust pressure. Typically,higher inlet pressures and higher inlet temperatures yield higher cycleefficiencies. Lower exhaust pressures typically also yield higher cycleefficiencies. Exhaust pressures are normally limited by ambient factors,such as the temperature of the river water, ambient air, or other fluidused to cool the condenser. This will set the limit for the exhaustpressure, and the condenser and associated equipment will be designed toapproach this limit, based upon evaluated parameters such as size,cooling medium available, environmental factors, and cost.

[0045] Design Limitations

[0046] Inlet pressure and inlet temperature are typically selected bythe plant design engineer. However, there are limits that are imposed inthese designs. As the inlet pressures are increased, the stresses on theboiler tubes, steam turbine casing, and steam turbine internals areincreased. These stresses impose limits on the manufacturer's ability toproduce this equipment, or economic limitations on the feasibility ofproducing this equipment. In addition, above 3206 psia, steam no longercan coexist as both water and steam. This point is referred to as thecritical point of steam, and above this pressure steam does not boil.Instead, both water and steam are a fluid and a more intricatesuper-critical boiler is required to produce steam above this pressure.At higher temperatures, the allowable stress of the boiler tubes, steamturbine casing, and steam turbine internals is reduced, and near thecurrent limits, conventional steam turbine materials rapidly loose theirproperties as the temperature is increased only small amounts (50° F.).Conventional large steam turbines built as state of the art machineshave HP inlet temperature limits in the range of 1050° F.

[0047] Steam Cycle Optimization

[0048] Once a boiler steam pressure and temperature is selected, thesteam cycle then must be optimized. A typical high efficiency steamcycle will involve the use of feedwater heaters, a reheater, a reheatsteam turbine, boiler feed pumps, and a condenser. A descriptivedocument on cycle optimization is an informative paper issued by GeneralElectric Company (GE) entitled “Steam Turbine Cycle Optimization,Evaluation, and Performance Testing Considerations” (General ElectricReference GER-3642E, 1996) by James S. Wright. This document providesrelative performance variations for different cycle parameters such aspressure, temperature, number of reheats, and number of feedwaterheaters.

[0049] Rankine Cycle Example

[0050]FIG. 3 is a schematic representation of a Rankine Cycle with bothfeedwater heating and reheat. This sub-critical Rankine Cycle works byproviding water to the inlet of the boiler feed pump (BFP) (301). Thewater is then pumped to a desired discharge pressure by the BFP (301).This pressurized water is then sent through the feedwater line (302) tofeedwater heater (FWH) (303) and through line (304) to feedwater heater(305). The feedwater heaters (303, 305) preheat the feedwater before itenters the boiler at the boiler inlet (306). This preheated feedwatertravels to the evaporator section (307) of the boiler where heat isadded to the pressurized water.

[0051] Steam exits the boiler section at (308) and continues tosuperheater section (309) and exits at (310). This superheated steam issent to the high-pressure (HP) section of the steam turbine (311). Thesteam expands through the HP section to (312), and then returns to thereheat section of the boiler (RHT) (313) where heat is added to returnthe steam typically to a temperature at or near the inlet steamtemperature. This reheat steam is then sent to the Intermediate Pressure(IP) section of the steam turbine at (314). This steam then expandsthrough the IP turbine section (315) and produces shaft horsepower. Thesteam then exits the IP section and via the crossover pipe (316) andgoes to the LP section of the steam turbine (317).

[0052] Due to the high volume flows at low-pressure, the LP section istypically a double flow section on large units, so steam enters themiddle of the casing and travels both forward and aft through theblading to produce more shaft horsepower. The steam then exhausts at(318) into the condenser (COND) (319). Condensed steam leaves thehotwell (330) and returns via the feedwater line (320) to the inlet ofthe BFP (301). For feedwater heating, steam is extracted from the IP andLP sections of the steam turbine at (321) and (324) and sent tofeedwater heaters (305) and (303) respectively via lines (323) and(326). Non-return valves are used in these lines, (322) and (325), toprevent backflow of steam to the ST in case of a trip (emergencyshutdown) condition when pressures in the turbine will rapidly drop tocondenser pressure. These valves are safety devices only, and are eitheropen or closed. Steam from these extraction lines preheats the feedwateron its way to the boiler. The steam from the extraction lines iscondensed in the feedwater heaters and the condensate (327, 328) isreturned to the inlet of the BFP (301). Again, shaft horsepower producedin the ST is converted into electrical power in the generator (GEN)(329).

[0053] For larger, central power plant applications, typical inletpressures for sub-critical applications are 1800 and 2400 pounds persquare inch gauge (psig). For supercritical applications, pressures of3500 psig and greater are employed. Inlet steam temperatures for mostlarge steam turbines are limited to about 1050° F. for both the inletand reheat steam. However, some advanced technology steam turbines areutilizing inlet temperatures of 1070° F. for the HP inlet and 1112° F.for reheat, as detailed in a descriptive document on steam turbinesissued by General Electric Company (GE) entitled “Steam Turbines forUltrasupercritical Power Plants” by Klaus M. Retzlaff and W. AnthonyRuegger (General Electric Reference GER-3945, 1996).

[0054] Rankine Cycle Efficiency Comparison

[0055] Based upon a steam turbine with a 90% efficiency, FIG. 4illustrates a relative comparison of a basic Rankine Cycle (Option 1),(excluding boiler efficiency and parasitic power requirements) to onethat uses only reheat (Option 2, Option 3), and to one that uses bothreheat and feedwater heating (Option 4, Option 5). Variations in theinlet pressure with reheat (Option 3) and feedwater heating (Option 5)are also included. Option 6 and Option 7 are for supercritical steamapplications. Option 6 is a supercritical steam cycle withultrasupercritical (inlet or reheat temperatures above 1050° F.) steamconditions and double reheat (steam is reheated twice, at two separatepressure levels, in the boiler). Option 7 is the same as Option 6 withthe addition of feedwater heating. For the purposes of this comparison,only two extractions were utilized and the extraction pressures wereassumed to be at the cold reheat pressure and the crossover pressure(2nd cold reheat for supercritical applications). More feedwater heaterswill yield even better cycle efficiencies. General Electric Company (GE)informative document entitled “Steam Turbine Cycle Optimization,Evaluation, and Performance Testing Considerations” (General ElectricReference GER-3642E, 1996) by James S. Wright provides data for theselection of the optimum number of feedwater heaters, stating that a1.5% heat rate penalty is assessed for only three feedwater heatersversus seven. Therefore, the feedwater heating cycle efficiency shown onFIG. 4 (Options 4, 5, and 7) has room for improvement. With reheat,optimum feedwater heating, and ultrasupercritical steam conditions,overall plant cycle efficiencies in excess of 45% are possible.

[0056] The overall plant cycle efficiency includes not only the basicsteam cycle efficiency as shown in FIG. 4, but also the boilerefficiency and parasitic power requirements such as the boiler feedpumps and the condenser circulating water pumps. As stated in POWERMAGAZINE, (ISSN 0032-5929, July/August 1998, page 26):

[0057] “Over the last few years, new designs have evolved to boostefficiencies of steam power plants, and the steam turbine is a largepart of this effort. Efficiencies of 45% (LHV) [Lower Heating Value] orhigher are now possible with the latest fossil-fired steam plants usingthe highest steam parameters, advanced feedwater heating cycles, boilerand turbine metallurgies, etc.”

[0058] To obtain an overall plant efficiency of 45% LHV, including theboiler efficiency and parasitic power requirements, typically means thatthe basic steam cycle efficiency must be even higher than 45%. With aboiler efficiency of 85%, parasitic power requirements of 2.5%, a ratioof HHV (higher heating value) to LHV (lower heating value) of fuel of1.11 (typical for natural gas), and a plant efficiency of 45% (LHV), thebasic steam cycle efficiency would calculate to

48.9%=0.45/(0.85×(1−0.025)×1.11)  (1)

[0059] As seen from FIG. 4, the use of a reheat steam cycle can increasethe basic Rankine Cycle efficiency by 4.79% at the tabulated steampressures. However, the use of reheat as well as increased inletpressures and feedwater heating can boost efficiency by at least 10.3%for sub-critical steam conditions. (Note that efficiency improvement isthe ratio of a particular option efficiency to the base efficiency.Thus, a 40% efficient cycle would convert 40% of the input energy toelectricity. That is twice as much as a 20% efficient cycle. Therefore,the efficiency improvement from a 20% efficient cycle to a 40% efficientcycle is 100%, or twice as much output).

[0060] Fuel efficiency is of the utmost importance at power plants and alarge central coal-fired power plant may expend approximately US$140million annually for fuel, assuming a plant rating of 1000 MW, 45%thermal efficiency LHV (lower heating value of the fuel), US$2.00 permillion BTU for fuel, and 8500 operating hours per year. Given thesefacts, even a 1% increase in efficiency will equate to large costsavings in fuel (US$1.4 million annually).

[0061] Combined Cycle Application

[0062] Although the Rankine Cycle has been well proven, today's morestrict energy and environmental standards require more emphasis beplaced on fuel efficiency and low emissions from power plants. As aresult, new combined cycle plants are being designed and built.

[0063]FIG. 5 is a conceptual schematic for a combined cycle application.In the general sense, combined cycle is not limited to a Brayton Cycletopping cycle and a Rankine Cycle bottoming cycle, but can be anycombination of cycles. The topping and bottoming cycles could be thesame cycle using different fluids. Either way, FIG. 5 would beapplicable. In FIG. 5, the topping cycle fluid (TCF) (501) enters thetopping cycle engine (TCE) (502) where fuel (CFT) (503) is added toraise its temperature. The fluid performs work that is converted by thetopping cycle engine into shaft horsepower. This shaft horsepower drivesthe topping cycle load (TCL) (504). This load could be an electricalgenerator, pump, compressor, or other device that requires shafthorsepower. The exhausted fluid from the topping cycle engine isdirected through an exhaust line (505) to a heat recovery device (HRD)(506), and then exhausts to an open reservoir (507).

[0064] For this example, the topping cycle is an open cycle. In otherwords, the topping cycle fluid is taken from a large reservoir anddischarges to that same reservoir. The heat recovery device (506)captures a portion of the topping cycle exhaust energy and transfers itto the bottoming cycle fluid (BCF) (508). In this example, the bottomingcycle fluid is heated at three separate pressure levels: a high-pressureline (509), intermediate pressure line (510), and low-pressure line(511). These fluids then travel to the bottoming cycle engine (BCE)(512) where it produces shaft horsepower to drive the bottoming cycleload (BCL) (513). Again, this load could be an electrical generator,pump, compressor, or other device that requires shaft horsepower.

[0065] From the bottoming cycle engine, the bottoming cycle fluid entersa heat exchanger (HEX) (514) where heat is rejected. The bottoming cyclefluid then enters a pump or compressor or other fluid transfer device(FTD) (515) where it is then returned to the heat recovery device (506).For this example, the bottoming cycle is a closed cycle, meaning thatthe bottoming cycle fluid is continuously circulated within a closedloop. There could be more than two cycles in this process, and any ofthe cycles could be either open or closed loop. This describes the basicfundamentals of a combined cycle application.

[0066] HRSG in Combined Cycles

[0067] In many cogeneration and combination GT/ST power plants builttoday, combined cycle plants have come to mean power plants that utilizea Brayton Cycle as the topping cycle and a Rankine Cycle as thebottoming cycle. These plants utilize a gas or combustion turbine (GT)as the prime mover (Brayton Cycle machine), with a boiler at the exhaustof the gas turbine to recover the waste heat. This boiler is typicallyreferred to as either a waste heat boiler (WHB) or a heat recovery steamgenerator (HRSG). It may also have burners in place to increase theexhaust gas temperature and produce more steam than that available fromjust the waste heat (supplemental firing). The HRSG produces steam thatis then sent to the steam turbine (ST) to produce more power. Due to thehigh temperatures of the working fluid in the GT (approximately 2400° F.for GE industry standard “F”-class technology machines and 2600° F. forWestinghouse industry standard “G”-class technology machines), andrecovery of the waste heat, the combined cycle plants are much more fuelefficient than the conventional steam plants. In addition, with advancesin GT technology and the use of either distillate oil or natural gasfuel, the emissions from the combined cycle plants are extremely low.FIG. 6 illustrates a typical combined cycle application.

[0068] The HRSG is distinctly different from a conventional RankineCycle boiler. A Rankine Cycle boiler is fueled by a variety of fuels,including oil, natural gas, coal, biomass, as well as others. TheseRankine Cycle boilers may also use a combination of fuels as well. TheHRSG may not utilize any fuels at all, but only capture and utilize theexhaust heat from the GT. If it is supplementary fired, the HRSG willrequire more refined fuels such as natural gas or distillate oil. Solidfuels such as coal and biomass are not typically utilized in these typesof boilers.

[0069] As seen from FIG. 6, there are numerous sections to the HRSG,including three evaporator sections (one for each pressure level),economizers, superheaters, and a reheater. Sections (601) and (602) areeconomizers. These are large tubed sections in the HRSG that preheatwater before it is converted into steam in the Evaporator. Sections(603), (606), and (609) are LP, IP, and HP evaporators respectively.Sections (604), (605), and (607) are feedwater heaters. Section (608) isthe IP superheater while sections (610) and (612) are HP superheaters.Section (611) is the reheater section. These HRSGs are typically verylarge and heavy pieces of equipment with literally miles of tubesinside.

[0070] Steam from each pressure level is utilized in the power plantwhere required, but essentially, most steam is generated for the purposeof producing additional power in the ST. This means that the lowerpressure levels of steam must be introduced or admitted to the ST at theproper point on the ST other than the HP inlet. It also means that theST must have provisions (openings, nozzles, connections, trip valves,etc.) where this steam may be admitted, and that at the operatingconditions the steam pressure in the ST at these connections must beless than the pressure of the steam from the HRSG corresponding boilersections. Otherwise, steam will not flow into the ST.

[0071] As noticed from a comparison of FIG. 6 with FIG. 3, theconventional Rankine Cycle utilizes feedwater heaters that take steamfrom the ST to preheat feedwater, while the HRSG utilizes the GT exhaustheat to provide this function. Therefore, conventional steam fedfeedwater heaters are not typically employed in combined cycleapplications. In GE informative document GER-3582E (1996), entitled“Steam Turbines for STAG™ Combined Cycle Power Systems”, M. Bossconfirms that feedwater heaters are not utilized in the prior art:

[0072] “Exhaust sizing considerations are critical for any steamturbine, but particularly so for combined-cycle applications. There areusually no extractions from the steam turbine, since feedwater heatingis generally accomplished within the HRSG”.

[0073] Another modification typically used for combined cycleapplications is the use of two boiler feed pumps (630), and (631),typically referred to as the LP and HP BFPs respectively. Thisarrangement allows the LP pump to provide pressurized water for the LPand IP pressure levels and the HP pump provides water for the HPpressure level, which saves pump horsepower. For large combined cycleapplications, the steam turbine/condenser arrangement is similar to theRankine Cycle depicted in FIG. 3, (although internally, the steam pathdesigns are totally dissimilar).

[0074] HRSG/Combined Cycle Disadvantages

[0075] General Disadvantages

[0076] With current technology, maximum inlet pressures to the steamturbine for combined cycle applications are nominally 1800 psia withinlet steam temperatures near the limit of 1050° F. for both the inletand reheat steam. Some of the disadvantages of this HRSG arrangement forcombined cycle applications are as follows:

[0077] 1. Steam cycle efficiencies are much lower than those ofconventional steam power plants.

[0078] 2. Multiple evaporator sections are required to maximize heatrecovery. This results in increased equipment and maintenance costs.

[0079] 3. Multiple evaporator sections require the plant operators andcontrol systems to monitor and control all boiler (evaporator) drumlevels.

[0080] 4. The HRSGs with the multiple sections are very large, requirelarge amounts of infrastructure building volume, large amounts of floorspace, and large foundations to support the weight of the HRSG.

[0081] 5. The HRSGs are expensive (approximately $10 million for a HRSGthat recovers exhaust gas heat from one GE Frame 7 GT).

[0082] 6. Maintenance increases with the number of components,evaporator sections, controls, and other devices.

[0083] 7. Low-pressure steam (steam other than the highest pressuresteam) has much less ability to produce power in the ST than higherpressure steam.

[0084] 8. Partial load, off design operation, and other conditionsbesides the design conditions typically have reduced heat recovery andlower cycle efficiencies.

[0085] 9. Increased amounts of tubing in the HRSG to enhance heatrecovery add flow restriction to the exhaust gases from the GT and thisincreased back pressure decreases GT output and efficiency.

[0086] 10. Gas turbine exhaust temperatures are not sufficient toproduce some of the elevated steam conditions now used in advanced steamcycles (600° C. which is equivalent to 1112° F.).

[0087] 11. Balancing problems in the reheat lines with multiple GTs(typically three or more) make it difficult to utilize large STs incombined cycle power plants in the prior art. For modem, large, andefficient combined cycle plants such as a GE S207FA, the steam turbinerating is approximately 190 MW, which is much smaller than GE's largesteam turbines which can exceed 1200 MW. For more information on largesteam turbines, reference the informative paper issued by GeneralElectric Company (GE) entitled “Steam Turbines for Large PowerApplications” by John K. Reinker and Paul B. Mason (General ElectricReference GER-3646D, 1996).

[0088] Part Load Operation Inefficiencies

[0089] Another disadvantage of the combined cycle application is partialload (part load) operation. As the system to which a power plant isconnected reduces its load requirement, the power plant must respond byproviding less output. This load modulation allows for a constant speedon the machinery and a constant frequency of power (e.g., 60 Hz in theUnited States and 50 Hz in Europe). To modulate the load at a combinedcycle plant, less fuel is burned in the GT, and the power output isreduced. This typically requires a reduction in the GT firingtemperature and/or a reduction in GT airflow.

[0090] Part load operation reduces the efficiency of the GT, thusreducing the efficiency of the entire combined cycle plant. FIG. 7illustrates a typical curve for a large modem GT with inlet guide vanes(IGVs) to modulate inlet airflow. Even with the enhanced part loadefficiency gained by the use of IGVs, at 60% load (GeneratorOutput—Percent Design), the GT consumes over 70% of the fuel required atfull load (Heat Consumption—Percent Design). This represents a 17.5%increase in heat rate (specific fuel consumption). For GTs without IGVs,this decay in performance would be even more pronounced.

[0091] To help offset this part load decay, plus provide more poweroutput for a given amount of hardware (sometimes referred to as powerdensity), manufacturers can provide combined cycle power plants with twoGTs, each with its own HRSG, feeding into one ST (referred to as a2-on-I arrangement). With an arrangement such as this, when the powerplant load decreases to slightly less than 50% for a 2-on-1 arrangement(2-GTs, 1-ST), one GT can be shut down, and the remaining GT can returnto near 100% output. This mode of operation increases part loadefficiency below 50% of total plant load as illustrated graphically inFIG. 8. This graphically illustrates a typical two GT comparison takenfrom GE informative document GER-3574F (1996), entitled “GECombined-Cycle Product Line and Performance” by David L. Chase, Leroy O.Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak for acurve of GE combined cycle part load performance with a 2-on-1arrangement. For a 3-on-1 arrangement, switchover from three to two GTscould occur slightly below 67% load. This still provides for substantialincrease in plant heat rate at part load conditions. Note that providingthis increase in part load efficiency occurs as a result of higherequipment costs. The prior art has yet to solve the efficiency problemwithout the addition of more equipment that increases the overall powerplant costs.

[0092] Supplementary Firing of HRSG

[0093] Another solution to add flexibility to the operation of acombined cycle power plant is the use of supplementary firing in theHRSG. This mode of operation is when fuel is burned in the HRSG justafter the GT (or at some intermediate point within the HRSG). Thisincreases the temperature of the exhaust gas to the HRSG and producesmore steam that can be sent to the ST. This allows the plant to producemore power. However, the plant heat rate increases, and fuel efficiencydecreases accordingly. This result is stated by Moore of GE in U.S. Pat.No. 5,649,416. This patent, as well as U.S. Pat. No. 5,428,950 byTomlinson, is referenced by Rice in U.S. Pat. No. 5,628,183. Therefore,supplementary firing of the HRSG is considered by the manufacturers tobe a means to obtain more output, but with a penalty on efficiency. GEinformative document GER-3574F (1996) entitled “GE Combined-CycleProduct Line and Performance” by David L. Chase, Leroy O. Tomlinson,Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak states that

[0094] “incremental efficiency for power produced by supplemental firingis in the 34-36% range based upon lower heating value (LHV) of thefuel.”

[0095] Also in this GE document, Table 14 indicates that HRSGsupplemental firing can increase combined cycle plant output in theprior art by 28%, but only with an increase in overall combined cycleheat rate (specific fuel consumption) of 9%. No technique has been shownin the prior art to eliminate this heat rate penalty associated withsupplemental firing.

[0096] Additionally, supplemental firing in the prior art can beutilized to achieve higher ST/GT ratios than is typical for conventionalcombined cycles. However, operation at these high levels of ST/GT outputare typically short in duration to meet peak power demands, and longterm operation at these ratios is not economical. Therefore,conventional combined cycle power plants that are designed with ST/GTratios approaching unity do not operate predominantly as Rankine cyclepower plants, but do so only to satisfy temporary peak plant loads, anddo so with a significant efficiency penalty at all operating conditions.

[0097] Gas Turbine Performance Decay

[0098] As mentioned in the discussion on the Brayton cycle,approximately ⅔ of the energy produced by the turbine section of the gasturbine is required to drive the compressor section, with the remaining⅓ available to drive a load. This power consumed by the compressor at67% of the turbine output, is much higher than the Rankine cycle examplewhere the boiler feed pumps (BFP) only consumed 2% of the turbine power.Therefore, the GT is susceptible to performance decay if the compressordoes not maintain optimum efficiency.

[0099] For example, a typical efficiency for an axial flow aircompressor used with a large GT might be 90%. Therefore, if thecompressor requires 67% of the turbine section output, the ideal power(100% efficient) would only be (0.67*0.90)=0.603 or 60.3%. If thecompressor efficiency were to decay by 2.5%, its new efficiency would be(0.90*0.975)=0.8775 or 87.75%. The compressor power required would nowbe (0.603/0.8775)=0.6872 or 68.72%. Turbine net output would be reducedfrom 33% (1.00-0.67) to 0.3128% (1.00-0.6872). This represents a 5.2%loss in output (0.3128/0.33=0.9479). Therefore, it can be readily seenthat small decreases in efficiency for the GT compressor lead to largedecreases in efficiency and output for a GT.

[0100] The efficiency and rating loss of 5% from the above example istypical of many GTs after about one or two years of operation. Thisefficiency decay is largely a result of worn clearances in thecompressor and erosion of the compressor blade tips. New blades andseals will typically restore the compressor efficiency to almost “new”condition efficiency. However, this is a costly and time consumingrepair, and would probably only be done at major inspections, which arescheduled approximately every four years for modem GTs. Therefore, plantowners and operators will need to plan on this performance decay betweenmajor overhauls of the GTs.

[0101] Candidates for Improvement in the Prior Art

[0102] From the foregoing discussion it can be seen that parameters ofthe current and defined technology that are candidates for improvementmay be described as follows:

[0103] Flexibility

[0104] Due to the electrical load demand in a particular region ormarketplace, the electric utility (which distributes electrical power tothe end users) determines the need for power based on current demand andfuture projections. For example, if this load was determined to be 850MW, in a conventional Rankine cycle configuration the utility/PowerDeveloper would contract with an Architect/Engineering (AE) firm todesign and build such a plant. The boiler, pumps, condenser, steamturbine, and all the other plant auxiliaries would then be designed forthe specified output of 850 MW. This can be accomplished largely due tothe fact that steam turbines are custom designed and manufactured.However, with gas turbines being production line items, and combinedcycles being primarily gas turbine based power plants, to achieve thehighest efficiencies and best capital cost, a utility and/or powerdeveloper can no longer specify just their plant output, but must findthe best fit for their needs from the available combined cycle offeringsfrom the various manufacturers. For example, a review of the availablecombined cycle plants from the 1997 TURBOMACHINERY HANDBOOK, (USPS871-500, ISSN 0149-4147), indicates that there are no 850 MW combinedcycle plants available for 60 HZ applications. Thus, a plant developer'sdesign flexibility is constrained by the current state of the art ofcombined cycle power plant equipment. This implies that in certaincircumstances the equipment complement for a given power plantinstallation will not be optimal because of constraints placed on plantequipment configurations by the current state of the art.

[0105] Efficiency

[0106] Combined cycle power plants are extremely energy efficientcompared to other conventional means of producing electricity. However,a large central combined cycle power plant rated for 1000 MW at 55%thermal efficiency LHV (lower heating value of the fuel) operating 8500hours per year at full load with a fuel cost of US$3.00 per million BTUof fuel will expend approximately US$175 million annually for fuel. Evena 1% increase in efficiency will equate to large savings in fuel(US$1.75 million annually).

[0107] In U.S. Pat. No. 4,333,310 issued to Robert Uram, a controlmethod is utilized which monitors the steam temperature to the ST andmodulates the afterburner (supplemental firing) to control thetemperature of the superheated steam. While providing optimum ST inlettemperatures, this function does little to affect load. In this patent,Uram states

[0108] “It is desired that the steam turbine be operated in what iscalled a “turbine following” mode wherein the plant is supplyingelectrical power to a load, such that the steam turbine follows the gasturbines and each afterburner positively follows a respective gasturbine. In other words, the heat contributed by the afterburner followsthe temperature of the gas turbine exhaust gas, and the steam producedby the gases exhausted from the afterburners is used in total by thesteam turbine.”

[0109] These teachings of the prior art are in direct contrast to thatof the present invention in which the heat contribution via supplementalfiring is independent of the gas turbines, and the gas turbines aredesigned to operate substantially at their optimal full rated capacity.

[0110] Installed Cost

[0111] Next to fuel costs, the largest cost for a combined cycle plantis typically debt service. Manufacturers, engineering firms, and ownersare always interested in finding ways to reduce the installed cost ofpower plants. At 8% interest and US$450 per kW of capacity, a 1000 MWcombined cycle power plant would have a debt service of approximatelyUS$45 million per annum for 20 years. Reducing the capacity cost, inUS$/kW, directly reduces the debt service.

[0112] Temporary Capacity Extension During Peak Demand Loading

[0113] One dilemma that faces power plant owners and utilities is theproper selection of power plant capacity. Selecting a plant that is toosmall results in power shortages, brownouts, and/or the need to purchaseexpensive power from other producers. Selecting a plant that is toolarge results in operation at lower efficiency during part load andincreased capital cost per kWh produced. In many situations the problemfaced by power plant developers is the need to provide for peak powerneeds and temporary demand loading. This peak may occur only in certainseasons for a limited span of time. Typically in the summer monthsduring peak hours on the hottest days is the most challenging time forpower producers to meet the system load. Having the ability to provideexcess capacity during this time period is highly desirable, and in theemerging arena of electrical power deregulation, it may prove to be verylucrative. For example, in the early summer of 1999, power shortages inthe Northeast United States have caused concern for the system's abilityto meet peak power demands. Some local newscasts have reported costs forcapacity at $30/MWh during normal periods and as high as $500/MWh duringpeak. However, even much greater capacity costs have been incurred, asreported in POWER MAGAZINE, (ISSN 0032-5929, March/April 1999, page 14):“Reserve margins are down nationwide from 27% in 1992 to 12% in 1998,according to Edison Electric Institute, Washington, D.C., becausederegulation uncertainty has caused capacity additions to stall. Lastsummer's Midwest [United States] price spikes, up to $7000/MWh, garneredmost of the press coverage, but spikes of $6000/MWh also occurred inAlberta . . . ”

[0114] However, providing peak power will not be lucrative if the powerplant owners have to pay for this capacity, pay the debt service, andyet make revenue on this extra capacity only during a few days of theyear. Therefore, power plants that can provide more output than normalduring peak demand hours are needed to help supply system load duringthese peak demands.

[0115] Reference FIG. 31B for a graphic illustrating the relativepercentage of time that a typical power plant spends in peak,intermediate, and base loading conditions. From this graphic it can besurmised that it would never be profitable to design a power plant topeak loading conditions, as they occur less than 10% of the time. Sinceprior art power plants are generally incapable of wide variations inpeak power output, the only practical option available for present powerproviders is to purchase power over the electrical grid during times ofpeak power demand. The present invention teaches a system and methodwhich permits this peak demand to be satisfied without the need forpurchasing external power over the electrical grid, thus providing aneconomic advantage over the prior art.

[0116] Non-Local Power Generation/Distribution Reliability Issues

[0117] One significant problem with the prior art is that the plantcapacity is in general a relatively fixed and narrow range of powergeneration operation. When peak power demands are placed on theelectrical grid, electrical power must be purchased from elsewhere onthe grid where electrical demand relative to remote plant capacity islower. There are several major problems with this mode of providing forpeak power by rerouting remotely generated power plant capacity.

[0118] First, there exist losses associated with transmission of powerfrom remote sites to the place where the electrical power is beingdemanded. For example, a hot summer day in New York City may requirediversion of power from Canada or the western United States, resultingin significant line losses during transmission.

[0119] Second, there is a reliability drawback in purchasing power fromdistant parts of the grid during periods of peak load. While it ispossible to redistribute power, the tradeoff is instability in theelectrical grid. What can happen is that small failures in remote partsof the grid can cascade throughout the grid to either cause additionalequipment failures or cause instability in the grid voltage. Thus, whilepurchasing power from remote power plants may alleviate some localreliability problems with respect to providing electric power, thetradeoff is an overall reduction in the reliability of the entireelectrical grid. Thus, relatively insignificant events in remote partsof the country can cascade throughout the electrical grid and result inserious electrical failures in major metropolitan areas.

[0120] Thus, given the above reliability concerns, it is in generalalways better to be able to provide electrical power local to the demandfor that power. While the existing prior art relies heavily on powersharing and distribution, the present invention opts for the morereliable method of generating the power locally to provide a powergeneration system that is more efficient and reliable that the currentstate of the art. It is significant to note that the prior artlimitations on plant output during peak load generally preclude localgeneration of the required peak power demand. This forces traditionalpower plants to purchase power from remote power plants at a substantial(10× to 250×) price penalty.

[0121] Operation and Maintenance Costs

[0122] Costs for personnel, fuel, maintenance, water, chemicals, spareparts, and other consumables, including other costs such as taxes andinsurance, all contribute to Operation and Maintenance (O&M) costs. Asthe plant size grows, the amount of equipment increases, and as thecomplexity of the equipment increases, O&M costs also increase. In thequest for higher efficiency, more elaborate and expensive technology isbeing utilized in the gas turbines. The maintenance costs associatedwith exotic new materials, intricate blades, and complex hardware isprojected to be significantly more expensive than the slightly lessefficient, proven gas turbine hardware and associated plant designs.

[0123] To be prepared for an equipment failure, plant owners must retainlarge quantities of spares on hand at their facility. This constitutesinventory that has high costs in terms of both unused capital and taxes.Methods to reduce O&M costs are always desired by the plant owners andoperators.

[0124] Fuel Gas Compression

[0125] Current projections are that natural gas will have a stablesupply and price structure until the year 2010. This fuel is clean,efficient, and inexpensive, and thus is the preferred fuel for combinedcycle applications. However, if the power plants are not located inclose proximity to major natural gas pipelines, the lower pressurenatural gas may have to be compressed to a sufficient pressure to beused in the GT. In addition, the higher efficiency GTs such as theWestinghouse model 501G require higher fuel gas pressure than GTs withlower pressure ratios, such as a GE model PG7241FA GT. This need forhigher pressure natural gas requires expensive natural gas compressorsthat are critical service items (the plant cannot operate without them).These natural gas compressors require frequent maintenance and alsoconsume parasitic power (the power to run the compressors reduces thenet power available from the power plant to the grid). Reducing the needfor these components reduces the plant installed cost, reduces realestate requirements, improves reliability, and increases the plant netoutput.

[0126] Plant Reliability

[0127] Electrical power reliability has become a facet that is demandedby both the residential consumer and industrial user of electricity.Therefore, the technology to produce power must be proven and reliable.In U.S. Pat. No. 5,628,183, Rice proposes a higher efficiency combinedcycle power plant. However, this system requires the use of diverters inthe HRSG, natural gas reformers, and the use of steam superheated to1400° F. These systems will all add greatly to the installed cost andO&M costs. In addition, to date, boiler tubes, HRSGs and STs have notdemonstrated long term reliable operation at elevated temperatures above1150° F., and HRSGs with diverters and natural gas reformers are as yetunproven in the marketplace.

[0128] Air Consumption

[0129] GT engines consume large quantities of air. A typical combinedcycle installation will consume approximately 20 lbs. of air per kW ofelectricity produced. This equates to approximately 260 cubic feet (atsea level) per kW. This air must be filtered before it enters the GT toprevent foreign object damage in the GT. Periodically, the air filtersmust be cleaned and/or replaced. This adds to the O&M costs andincreases plant downtime (time when the plant is out of service andunavailable to produce power).

[0130] In addition, the air consumed by the GT is discharged to the HRSGand then exhausted to atmosphere. As more air is consumed, more air mustbe exhausted. This represents an efficiency loss as the HRSG exhausttemperature is typically about 180° F. In addition, this airflow servesto heat the atmosphere and contribute to local air quality problems.

[0131] Plant Emissions

[0132] In order to obtain a permit to operate, a power plant must firstobtain an air permit. This permit typically states the allowable levelsof certain criteria pollutants that a plant may emit. Combined cyclepower plants are very clean producers of power compared to otherconventional methods, but are typically plagued by one criteriapollutant, nitrous oxides (NOX). This criteria pollutant is usuallycontrolled by steam and/or water injection into the GT, dry low NOXcombustion systems, and/or exhaust gas aftertreatment. The exhaust gasaftertreatment typically employed is “Selective Catalytic Reduction”(SCR) which essentially works by injecting ammonia (NH.sub.3) into theexhaust gas stream in the presence of a catalyst at a specifiedtemperature range to return the NOX formed by the combustion processinto N2 and H₂O.

[0133] In U.S. Pat. Nos. 3,879,616 by Baker, et al., 4,578,944 byMartens et al., and 5,269,130 by Finckh, et al., the plant load iscontrolled by changes in the GT output. However, at partial load, GT NOXemissions are typically increased. Therefore, it may be necessary tointroduce more ammonia into the exhaust gases for emission reduction.This increases O&M costs, and can be significant to the point where, atthe plant design stage, the desired GTs cannot be used due to highemission levels at part load operation. Also, if run at full load, someplants may not require SCR, but due to part load operation, SCR will berequired. Another factor related to emissions is air consumption. GTsrequire large amounts of air, and the more air that is consumed, themore potential there is for emissions.

[0134] Environmental Considerations

[0135] Besides air emissions, a power plant must be concerned with otherenvironmental impacts as well. To operate a steam plant, a clean sourceof water must be available to provide make-up water. This make-up wateris used to replace steam/water that is lost to ambient through leaks,blowdown, or other loss. Blowdown is the water that is taken from theevaporator sections of the HRSG and dumped to the sewer. This blowdowntypically is taken from a low point on the HRSG to remove feedwater thathas high concentrations of minerals and deposits. This process helpskeep the steam path clean and minimizes ST deposits and blade failuredue to stress corrosion cracking. This blowdown must be discharged intorivers, streams, etc. and as such requires water permits that may bedifficult and time consuming to obtain from regulatory authorities.

[0136] Distributed Plant Control System (DCS)

[0137] Modern combined cycle plants typically use a distributed controlsystem (DCS) to control the entire plant. These DCS controls integratewith the individual control systems on the GTs and STs. Many otherparameters can be monitored and controlled by the DCS. Use of controlsto better either efficiency or operation is described in U.S. Pat. No.3,879,616 by Baker et al., U.S. Pat. No. 4,201,924 by Uram, and U.S.Pat. No. 4,578,944 by Martens et al. None of these patents, however,provide control of heat transfer in the HRSG. In U.S. Pat. No. 5,269,130by Finckh et al., a method of controlling excess heat in the HRSG isutilized for part load operation of the GT. This method, however, doesnot provide comprehensive control, but only a means for recovering lowtemperature waste heat. None of the aforementioned patents has devised amethod to control the exhaust gas temperature of the HRSG to its optimumtemperature.

[0138] Plant Operational Efficiency

[0139] Combined cycle power plants in the prior art that are designedfor maximum efficiency typically utilize multi-pressure HRSGs, commonlyat three pressure levels. For each HRSG, and for each pressure level,the operations staff must monitor the steam drum level. Also, parameterssuch as water quality and chemical content must be monitored for eachHRSG. Since the system load for any utility is constantly changing,combined cycle power plants are required, like other power producingplants, to be dispatched, or provide load as required to the electricalgrid. This means the power plant will not operate at a fixed load, butwill constantly be modulating load to meet the system demand. Toincrease load, supplementary firing (additional fuel burned at or nearthe inlet to the HRSG to add energy to the exhaust gases) can beaccomplished. However, this is detrimental to overall plant efficiency.This is noted by Rice in patent U.S. Pat. No. 5,628,183 with referencesto Westinghouse and General Electric studies. Moore in patent U.S. Pat.No. 5,649,416 states that

[0140] “Supplemental firing of the heat recovery steam generator canincrease total power output and the portion of the total power producedby the steam turbine, but only with a reduction in overall plant thermalefficiency.”

[0141] Therefore, it is common in combined cycle plants to see little orno supplemental firing used. Therefore, to change and meet varyingsystem loads, the GTs are brought from full load to part load operation.

[0142] As well as increasing emission levels as previously mentioned,this part load operation also has a detrimental effect on efficiency.FIG. 7 is a representative curve of GT efficiency versus load. At 100%load it consumes 100% fuel, however, at 60% load, it consumes 70.5% offull load fuel. This is an increase of 17.5% in specific fuelconsumption. For large central power plants, this factor equates tosignificant added fuel costs. In addition, operation at part load on theGT typically increases the emission levels for the most difficultcriteria pollutant, NOX. Part load operation of the GT also changes theexhaust gas flow through the HRSG. This change in flow upsets the heattransfer in the HRSG since this device is constructed with fixed heatexchange surface area. This phenomenon, as well as reduced GTefficiency, contributes to poorer overall efficiency at part loadoperation. If part load operation changes temperatures in the HRSGsignificantly, this could lead to ineffective operation of the SCR.

[0143] Steam Turbine Exhaust End Loading

[0144] Besides inlet pressure and temperature limitations, anothercommon limitation for the steam turbine (ST) is the exhaust end loading.This essentially is a function of two parameters, exhaust end flow andexhaust pressure. These two factors essentially determine the volumetricflow through the last stage blading of the ST. For optimum operation,there is a range of volumetric flow typically specified by the STmanufacturers. As this volumetric flow increases, larger blades and/ormore exhaust sections may be required.

[0145] However, due to mechanical limitations (centrifugal force), oncethe largest available blade volumetric limits are reached, more sectionsand more blades must be added to the exhaust end of the ST toaccommodate this flow. This adds to the installed cost and increases thereal estate requirements of the ST. Due to its configuration, aconventional combined cycle sends HP steam to the ST HP inlet, then addssteam from the IP section of the HRSG to this flow at the ST IP sectioninlet, then adds more steam from the LP section of the HRSG to this flowat the ST LP section inlet.

[0146] Therefore, in this arrangement, the HP and IP sections of the STsee relatively lower flows and lower volumetric efficiencies than the LPsection. This arrangement leads to STs that are at or near the exhaustend limit of the ST. This provides for little in the way of temporarycapacity extension for peak power production and leaves little or noability to uprate (increase) the ST in the future to a higher powerrating. Overall, this ST arrangement is less efficient than conventionalsteam plant STs since the HP and IP sections have low volumetric flows.

[0147] In GE informative document GER-3582E (1996), entitled “SteamTurbines for STAG™ Combined Cycle Power Systems”, by M. Boss, the authordiscusses the exhaust end loading that is associated with STs in theprior art:

[0148] “Exhaust sizing considerations are critical for any steamturbine, but particularly so for combined-cycle applications. There areusually no extractions from the steam turbine, since feedwater heatingis generally accomplished within the HRSG. Generation of steam atmultiple pressure levels (intermediate pressure and/or low-pressureadmissions to the turbine downstream of the throttle) increases the massflow as the steam expands through the turbine. Mass flow at the exhaustof a combined cycle unit in a three-pressure system can be as much as30% greater than the throttle flow. This is in direct contrast to mostunits with fired boilers, where exhaust flow is about 25% to 30% lessthan the throttle mass flow, because of extractions from the turbine formultiple stages of feedwater heating”.

[0149] Real Estate

[0150] A combined cycle installation, although typically smaller thanconventional steam plants, still occupies a large area. The HRSGs withtheir stacks are particularly large and require a great deal of floorarea (the HRSG for one Westinghouse model 501G gas turbine isapproximately 40 feet wide, 70 feet high, and 200 feet long). With thetrend towards deregulation of electrical power, plant owners will beseeking the ideal site for their power plants. In many instances, thisis near to the electrical load, which is usually in either an urban orindustrial area. This puts the power plant close to the end user ofelectricity, and eliminates the need for high voltage transmission lines(which also require large amounts of real estate). However, availablereal estate for a large combined cycle power plant may be difficult andexpense to attain in these areas.

[0151] Some prime real estate for these combined cycle power plants willbe existing power plants that can be repowered as combined cyclefacilities. These sites have the advantage of being properly zoned withthe necessary electrical and mechanical infrastructure. The drawback isthat the site may lack the necessary real estate for a combined cyclerepowering project. Therefore, it is desirable from a space efficiencyviewpoint as well as from a cost perspective to keep plants as small aspossible.

[0152] Noise/Public Acceptance

[0153] Public acceptance is becoming increasing difficult for manyutility power plant projects. Factors such as noise, traffic increase,unsightliness, pollution, hazardous waste concerns, and otherscontribute to public disapproval of power plants in close proximity topopulated areas. A plant that can be built smaller, quieter, with lessequipment, lower emissions, and maintain a low profile is preferred overa larger, more obvious plant. Therefore, more compact, higher “powerdensity” (power per unit volume) combined cycle power plants aredesired.

[0154] However, to meet the current trends in demand for powerconsumption, conventional power plants being constructed today simplyreplicate existing proven plant designs to meet the increased energyconsumption demand. No attention is currently being given to the issueof whether plants may be redesigned to consider the ancillary issuesassociated with the public acceptance of the plants themselves.

[0155] Heat Rejection

[0156] Both conventional steam and combined cycle power plants requiresome form of heat rejection. This is typically to condense thelow-pressure steam from the ST exhaust back into water. This heatrejection can be to the air, river, lake, or other “reservoir” that willabsorb the heat. Since this heat rejection will have an effect on thelocal environment and possibly the local biological life (i.e., fish ina river), methods to reduce heat rejection requirements are always indemand.

[0157] Gas Turbine Performance Decay

[0158] Although combined cycle power plants demonstrate highefficiencies, these efficiencies are for “new” power plants. Since thecombined cycles in the prior art are primarily GT based, theirefficiency levels are very susceptible to GT performance decay, aphenomenon in which the efficiency of the GT degrades substantially (2%to 6%) within only a year or two of operation. This can be a significantfactor in the cost of fuel as the overall combined cycle efficiency alsodegrades as the GT performance decays.

OBJECTS OF THE INVENTION

[0159] Accordingly, the objects of the present invention are tocircumvent the deficiencies in the prior art and affect the followingobjectives:

[0160] 1. Provide a combined cycle power plant that has more designflexibility than current offerings so that developers can havestate-of-the-art facilities, but purchase them at the capacity theyneed.

[0161] 2. Reduce overall fuel consumption at rated output, butespecially at part load conditions, as the plant will likely spend onlya small fraction of its operating time at rated load.

[0162] 3. Reduce installed cost of the power plant such that the debtservice is substantially reduced and that financing by a bank or otherlending institution is much easier for the owner.

[0163] 4. Leverage the time value of money with regards to capital,maintenance, and fuel costs to make the creation of power plants moreeconomically efficient and hopefully reduce the overall cost of electricpower generation.

[0164] 5. Provide the ability for the power plant to meet peak demandloads without sacrificing normal operation efficiency or significantlyincreasing the installed cost.

[0165] 6. Reduce inefficiencies and losses associated with thetransmission of power over long distances.

[0166] 7. Increase the overall reliability of the electrical grid bypermitting electrical power to be generated local to the demand duringtimes of peak demand loads.

[0167] 8. Reduce O&M costs. Besides fuel costs, the objective is also toreduce costs for maintenance, supplies, inventory, insurance, and otheroperating expenses.

[0168] 9. Reduce the need for fuel gas compression.

[0169] 10. Improve reliability.

[0170] 11. Reduce air consumption and air filtering requirements.

[0171] 12. Lower emissions of criteria pollutants, especially NOX.

[0172] 13. Minimize the discharge of water from HRSG blowdown and othersources.

[0173] 14. Utilize controls to the maximum extent feasible to increaseefficiency, reliability, and heat recovery.

[0174] 15. Simplify operation and devise methods and/or strategies toincrease part load efficiency and reduce emission levels.

[0175] 16. Optimize the ST efficiency by utilizing designs with improvedvolumetric efficiency and excess capacity to meet peak power demands.

[0176] 17. Conserve space and land mass required to house the powerplant by designing a compact, high power density arrangement.

[0177] 18. Reduce noise, size, space requirements, and equipment tominimize the effect the power plant has on local residents and thecommunity.

[0178] 19. Keep heat rejection to a minimum.

[0179] 20. Provide for economic and space efficient retrofit of existingsteam power plant and combined cycle installations so as to reducecapital costs and the economic burden associated with major equipmentadditions and added real estate requirements.

[0180] 21. Provide economic incentive for new plant construction to useenvironmentally friendly designs.

[0181] 22. Design combined cycle power plants that are less susceptibleto gas turbine performance decay.

[0182] These objectives are achieved by the disclosed invention that isdiscussed in the following sections.

BRIEF SUMMARY OF THE INVENTION

[0183] Briefly, the invention is a system and method permitting the useof fewer and/or smaller gas turbines (GTs) and heat recovery steamgenerators (HRSGs) in a combined cycle application. This conventionalcombined cycle equipment is replaced by a larger steam turbine andcontinuously fired heat recovery steam generators to provide a varietyof economic, energy conservation, and environmental benefits.

[0184] Present technology utilizes multi-pressure HRSGs to maximize theheat recovery from exhaust gases of a GT. This arrangement is commonlyused because the prior art teaches away from using continuously firedHRSGs because of the common belief that these configurations have lowerthermal efficiencies. Despite this commonly held belief, the presentinvention teaches that continuously fired HRSGs can be configured withthermal efficiencies on par or better than current combined cyclepractice. However, to obtain this level of efficiency, the continuouslyfired HRSGs and ST must be configured and designed differently thancurrent practice.

[0185] In several preferred embodiments of the present invention, theGTs are unchanged from the present art and exhaust to an HRSG. ThisHRSG, however, is designed as a single pressure level steam generator(SPLSG) (or primarily a single pressure level) which is optimized forcontinuous firing to produce higher pressure steam than in conventionalcombined cycle practice. In addition, the HRSG is designed to havecontrolled feedwater flows through the economizer/feedwater sections tomaximize heat recovery. Also, the ST is designed as a larger unit,typical of that which would be found in a conventional Rankine Cycleplant, with reheat and conventional ST extraction steam fed feedwaterheaters to maximize plant thermal efficiency. This benefit of a largerST typical of a conventional steam plant is described by Moore in patentU.S. Pat. No. 5,649,416 which is assigned to General Electric:

[0186] “Conventional steam power plants benefit in both lower cost andhigher efficiency through the economies of scale of large ratings. Atraditional rule of thumb regarding cost is that the doubling of plantrating results in a ten percent reduction in cost. The cost of one largegenerating unit according to this rule would be expected to cost on theorder of ten percent less than that for a plant with two half-sizeunits. Efficiency is also improved with increased size and powerratings. As with all turbomachinery, the internal efficiency of thesteam turbine is a strong function of the inlet volumetric flow, whichis directly proportional to the rating. Also, as is well known, thethermal efficiency of the Rankine Cycle increases with the pressure atwhich steam is generated. Increasing pressure, however, reduces thevolumetric flow of the steam at the turbine inlet, reducing the internalexpansion efficiency. The offsetting effect in overall efficiency,however, is much greater at low volumetric flow than at high volumetricflow. Therefore, an additional performance related benefit of increasingturbine size is that higher steam throttle pressure can be utilized moreeffectively.”

[0187] With the use of ample supplemental firing in the HRSG, thebottoming cycle with the present invention is given the liberty to bemore independent from the GT operation. Therefore, the GTs can beoperated at full load while the overall plant load is modulated over awide range of its full load capability by only changing the supplementalfiring rate and the STs load. This increases the overall plant ratingwhen utilizing a given set of GTs, provides flexibility for the combinedcycle plant rating through variation in the rate of supplemental firing,as well as increases the overall plant thermal efficiency at part load.In addition, it simplifies operation, and has the potential to reduceemissions.

[0188] By designing the HRSGs to be capable of firing to 2400° F., anexemplary single 2-on-1 arrangement of two GTs and one large ST replacestwo 2-on-1 arrangements (4-on-1 arrangements are typically not availablewhen reheat is utilized due to balancing problems on the reheat lines).This exemplary configuration saves two GTs, two HRSGs, one ST, threeswitchgear, three transformers, and the accessories, real estate, andmaintenance required to support this equipment. Capital costs for thepower plant in US$/kW are thus greatly reduced using the teachings ofthe present invention.

[0189] All this is accomplished by utilizing proven turbomachinerytechnology and hardware. The continuously fired HRSG with a singlepressure is a novel concept for this application, but is not beyondtechnological practice nor capability for implementation in the currentart. Therefore, there are little or no compromises in reliability. Thegeneral architecture for several preferred embodiments of the presentinvention is illustrated in FIG. 13, with several exemplary embodimentshaving more detail illustrated in FIG. 9 and FIG. 15.

[0190] Improvements Over the Prior Art

[0191] The present invention solves the problems present in the priorart by achieving the following objectives:

[0192] 1. Providing more design flexibility in the combined cycle powerplant so that developers can still achieve state-of-the-art efficiency,but yet specify the capacity they need.

[0193] 2. Reducing overall fuel consumption by improving both full loadand part load efficiency.

[0194] 3. Reducing installed costs by increasing the power density ofthe installation (more power output per a given amount of equipment).

[0195] 4. Reducing the overall cost of producing electricity by reducingthe three major factors associated with its production: fuelconsumption, capital costs, and maintenance costs.

[0196] 5. Provide temporary capacity for attaining peak loads byutilizing supplemental firing to produce more steam, as well as havingthe option to operate the ST at overpressure (inlet pressure slightlyabove rated) and reducing extraction steam flow to the feedwaterheaters.

[0197] 6. Increasing the efficiency of the power grid by permittinglocal generation of power during periods of peak loading. By permittinglocal power generation during these peak periods, inefficienciesassociated with “importing” power from other areas of a given country(and outside a country) are reduced or eliminated. (These are energylosses associated with transmitting power through power transmissionlines).

[0198] 7. Increasing the reliability of the electrical power grid byreducing the long haul transmission of electrical power during times ofpeak power loading.

[0199] 8. Reducing O&M costs, primarily by reducing the amount ofequipment and systems and utilizing equipment that has lower maintenancecosts per kWh produced (low maintenance cost STs versus high maintenancecost GTs).

[0200] 9. Minimizing the need for fuel gas compression by utilizingfewer GTs and GTs with lower fuel gas pressure requirements in the cyclein conjunction with a larger ST.

[0201] 10. Improving reliability by reducing the complexity of the powerplant design.

[0202] 11. Reducing air consumption by utilizing fewer GTs.

[0203] 12. Lowering emissions of criteria pollutants, especially NOX, byoperating the GTs at a steady, low emissions operating point, utilizingcleaner GTs, and utilizing fewer GTs.

[0204] 13. Minimizing blowdown and other discharge through higherefficiency cycles that require less steam flow per kW of electricitygenerated.

[0205] 14. Utilizing controls to increase efficiency, reliability, andheat recovery.

[0206] 15. Simplifying operation by running the GTs at full load over awide range of operation (total combined cycle plant output) and reducingHRSG pressure levels to only one.

[0207] 16. Maximizing ST efficiency by increasing volumetric flows,especially in the HP and IP sections.

[0208] 17. Conserving space and land mass with less equipment and higherpower density designs.

[0209] 18. Reducing noise, size, and space requirements with lessequipment.

[0210] 19. Keeping heat rejection to a minimum by utilizing highefficiency cycles with less heat rejection per kWh produced.

[0211] 20. Providing a combined cycle design that is more compatiblewith existing steam power plants allowing for more compact and costeffective retrofits of these existing plants to high efficiency combinedcycle technology.

[0212] 21. Minimizing air consumption, emissions of criteria pollutants,and heat rejection to the atmosphere, but providing these environmentalbenefits with lower cost than the conventional combined cycles.

[0213] 22. Reducing the impact of gas turbine performance decay byutilizing a combined cycle power plant that is less dependent upon thegas turbines and their efficiency.

BRIEF DESCRIPTION OF THE DRAWINGS

[0214] For a fuller understanding of the advantages provided by theinvention, reference should be made to the following detaileddescription together with the accompanying drawings wherein:

[0215]FIG. 1 illustrates a basic Rankine thermodynamic cycle;

[0216]FIG. 2 illustrates a schematic of a conventional prior art powergeneration system implementing the basic Rankine Cycle;

[0217]FIG. 3 illustrates a schematic of the Rankine Cycle including areheat cycle and extraction steam feedwater heating as applied to aconventional prior art power plant application;

[0218]FIG. 4 illustrates a comparison table of efficiencies between thebasic Rankine Cycle and the Rankine Cycle with various efficiencyenhancements;

[0219]FIG. 5 illustrates a schematic of the basic principles of acombined cycle;

[0220]FIG. 6 illustrates a schematic of the prior art for a combinedcycle power plant utilizing gas turbines, HRSGs, and steam turbines;

[0221]FIG. 7 illustrates a curve of heat consumption versus generatorpower output for an industry standard General Electric (GE) ModelPG7241(FA) Gas Turbine;

[0222]FIG. 8 illustrates part load performance for a General Electriccombined cycle power plant with two GE S207 GTs via graphs indicatingperformance characteristics for one and two gas turbine (GT) operation;

[0223]FIG. 9 illustrates a general arrangement of one preferredembodiment of the present invention as applied to the application of anelectric power plant;

[0224]FIG. 10 illustrates a tabular comparison of the efficiencies thatmay be realized using the teachings of the present invention as comparedto the prior art;

[0225]FIG. 11 illustrates a typical graph of steam enthalpy versustemperature at 1800 psia pressure assuming water as the motive fluid;

[0226]FIG. 12 illustrates a typical graph of gas turbine exhaust gasenthalpy versus exhaust gas temperature;

[0227]FIG. 13 illustrates a schematic of the general principles of thepresent invention as implemented in a combined cycle application;

[0228]FIG. 14 illustrates a typical graph of required log meantemperature difference (LMTD) versus fluid flow for a superheater andreheater application;

[0229]FIG. 15 illustrates an exemplary embodiment of a combined cyclepower plant application utilizing the teachings of the presentinvention;

[0230]FIG. 16 illustrates an exemplary system control flowchart that maybe used to control one or more heat recovery steam generators (HRSGs) asper the teachings of the present invention;

[0231]FIG. 17 illustrates an overall exemplary system control flowchartthat may be used to provide overall power plant system control as perthe teachings of the present invention;

[0232]FIG. 18 illustrates an exemplary system control flowchart whichmay be used to control and direct increased power plant output as perthe teachings of the present invention;

[0233]FIG. 19 illustrates an exemplary system control flowchart whichmay be used to control and direct decreased power plant output as perthe teachings of the present invention;

[0234]FIG. 20 illustrates an exemplary system control flowchart whichmay be used to control and direct transitional power control as per theteachings of the present invention;

[0235]FIG. 21 graphically illustrates the sources of energy inputs,losses, and efficiencies that are accounted for in an overall energyflow analysis;

[0236]FIG. 22 illustrates a typical GE 207FA combined cycle power plantconfiguration;

[0237]FIGS. 23A and 23B illustrate tabulated performance data for atypical GE 207FA 521 MW combined cycle power plant assuming a typicalprojected operation profile;

[0238]FIG. 24 illustrates a typical Westinghouse 2X1 501G 715 MWcombined cycle power plant configuration;

[0239]FIGS. 25A and 25B illustrate tabulated performance data for atypical Westinghouse 2X1 501G 715 MW combined cycle power plant assuminga typical projected operation profile;

[0240]FIG. 26 illustrates a typical 725 MW combined cycle power plantdefined by a preferred embodiment of the present invention;

[0241]FIGS. 27A and 27B illustrate tabulated performance data for a 725MW preferred embodiment of the present invention using a water-walledHRSG assuming a typical projected operation profile;

[0242]FIG. 28 graphically illustrates the relative part load performancedifference between a conventional combined cycle power plant and apreferred embodiment of the present invention;

[0243]FIG. 29 graphically illustrates several exemplary power plantconfigurations and their nominal range of available specified powerratings using the teachings of the present invention;

[0244]FIG. 30 graphically illustrates the basic steam cycle efficiencyrequired for an exemplary power plant configuration utilizing twoindustry standard General Electric (GE) Model PG7241(FA) Gas Turbines tomeet prior art efficiency levels over a range of power ratings;

[0245]FIG. 31A graphically illustrates a typical hourly regional systemload curve (from “Electricity Prices in a Competitive Environment:Marginal Cost Pricing of Generation Services and Financial Status ofElectric Utilities” (DOE Report number DOE/EIA-0614));

[0246]FIG. 31B graphically illustrates a typical load duration curvewhich depicts the overall long term use of rated plant capacity (dataobtained from Duke Energy Power Services, Inc.,http://www.panenergy.com/power/epdb2_(—)5.htm);

[0247]FIG. 32 illustrates a typical conservative weekly load profileutilizing the data contained in FIG. 31A;

[0248]FIG. 33 graphically illustrates the part load efficiencies ofseveral exemplary power plants of the present invention as well asseveral examples from the prior art;

[0249]FIG. 34 tabulates an economic comparison of an exemplary powerplant utilizing the teachings of the present invention to both a GES207FA combined cycle power plant and a Westinghouse 2X1 501G combinedcycle power plant, both from the prior art;

[0250]FIG. 35 is a typical heat balance process flow diagram for thesubcritical exemplary power plant embodiment of the present inventionused in FIGS. 26, 27A, 27B, 28, 33 and 34;

[0251]FIGS. 36, 37, and 38 tabulate some of the process data associatedwith FIG. 35;

[0252]FIG. 39 is a heat balance process flow diagram for theultrasupercritical exemplary power plant embodiment of the presentinvention used in FIG. 33;

[0253]FIGS. 40, 41, and 42 tabulate some of the process data associatedwith FIG. 39;

[0254]FIG. 43 graphically illustrates a power plant load control methodthat may be used with a combined cycle of the present invention in whichtwo or more GTs are utilized;

[0255]FIG. 44 tabulates data for the comparison of a retrofit of anexisting steam power plant to combined cycle technology between thepreferred embodiment and the prior art;

[0256]FIG. 45 illustrates a preferred embodiment combined cycle powerplant utilizing a hybrid fuel arrangement with a combustible fuel (CF)boiler;

[0257]FIG. 46 illustrates a preferred embodiment combined cycle powerplant utilizing a hybrid fuel arrangement with a nuclear reactor,geothermal steam generator, or other steam producing energy source;

[0258]FIG. 47 is an exemplary design/financing process flowchart fordetermining a preferred and/or optimal arrangement of a given inventionembodiment for a particular power plant application;

[0259]FIG. 48 is an exemplary plant economics process flowchart fordetermining a preferred and/or optimal arrangement of a given inventionembodiment for a particular power plant application;

[0260]FIG. 49 is an exemplary plant retrofit process flowchart fordetermining a preferred and/or optimal arrangement of a given inventionembodiment for a particular power plant retrofit application;

[0261]FIG. 50 is an exemplary hybrid fuel design process flowchart fordetermining a preferred and/or optimal arrangement of a given inventionembodiment for a particular power plant application utilizing hybridfuel;

[0262]FIG. 51 illustrates a GE three casing, four-flow steam turbinewith a combined HP/IP section and two double flow LP sections.

DESCRIPTION OF THE PRESENTLY PREFERRED EXEMPLARY EMBODIMENTS

[0263] Exemplary Disclosure

[0264] While this invention is susceptible to embodiment in manydifferent forms, there is shown in the drawings and will herein bedescribed in various detailed preferred embodiments of the inventionwith the understanding that the present disclosure is to be consideredas an exemplification of the principles of the invention and is notintended to limit the broad aspect of the invention to the embodimentillustrated.

[0265] Diagrams and Flowcharts

[0266] It should be noted specifically within the context of thedescriptions given in this document that schematics, flowcharts,diagrams, and the like may be augmented with components and/or stepswith no reduction in the generality of the teachings of the presentinvention. Similarly, components and/or steps may be removed and/orrearranged in the following descriptions with no loss of generality.This notice is especially important with respect to exemplary processflowcharts, in which the teachings may be used by one skilled in thecomputer arts to generate control systems that are functionallyequivalent, but which may rearrange or modify the disclosed steps andprocesses yet achieve the results as dictated by the present inventionteachings.

[0267] Equipment

[0268] Throughout the discussion of the present invention containedthroughout this document mention will be made to specific equipment fromGeneral Electric, Westinghouse, and other manufacturers. Specifically,much of the disclosure makes reference to the GE model S207FA powerplant comprising GE model PG7241FA gas turbines as well as comparableequipment by Westinghouse and others. These references are exemplaryonly, and given to provide the reader who is skilled in the art aframework in which to understand the teachings of the present invention.

[0269] Rather than speak in terms of fictitious equipment which may notbe familiar to those skilled in the art, this disclosure attempts to bemore practical by illustrating the teachings of the present invention interms of equipment that one skilled in the art will be familiar with andwhich is currently in use within the electric power industry. Nothing inthis disclosure should be interpreted to limit the scope of theteachings of the present invention to a specific manufacturer or modelof equipment. On the contrary, the present disclosure should beinterpreted as broadly as possible with respect to the equipment towhich the teachings may apply.

[0270] Overview

[0271] Steam has been used for power applications for decades, datingback to steam locomotives that burned solid fuel such as wood or coal toproduce power. Up to and into the 1980's, steam power plants were stillproducing the bulk of the electrical power in the United States ineither coal, oil, or nuclear-fueled power plants.

[0272] However, by the 1980's, many smaller cogeneration power plantswere being designed and built. These plants utilized a gas turbine astheir main engine with a heat recovery steam generator (HRSG) connectedto the exhaust of the gas turbine to recover waste heat (typically 900°F. to 1200° F. exhaust gases) and convert it into steam. This steam wasthen utilized for various purposes, district heating, process steam, orgeneration of additional power in a steam turbine. This plantconfiguration, gas turbine, HRSG, and steam turbine became known as acombined cycle arrangement, and due to its high efficiency, low cost,and ease of construction, has become the preferred power plant for theemerging Independent Power Producers (IPPs).

[0273] However, through evolution, this combined cycle power plant hasbecome a power plant that utilizes the gas turbine as its prime engineand the steam turbine as its secondary engine. An examination of thestandard combined cycle packages offered by gas turbine manufacturerstoday will verify this statement, as in most combined cycle plants inthe prior art, the gas turbines produce about two thirds of the totalpower output, with the steam turbines producing about the remaining onethird. A review of the manufacturer's standard combined cycle offeringswill illustrate this trend. The 1997 TURBOMACHINERY HANDBOOK, (USPS871-500, ISSN 0149-4147), tabulates standard combined cycle power plantsavailable from various manufacturer's including ABB, General Electric,and Westinghouse. In most every instance, the steam turbine(s) output iswithin the range of 40% to 60% of the gas turbine(s) output. GeneralElectric informative document GER-3567G, 1996, “GE Heavy-Duty GasTurbine Performance Characteristics”, by Frank J. Brooks, provides theoutput for the gas turbines used in their combined cycle power plants.

[0274] Several preferred embodiments of the present invention recognizethe combined cycle arrangement for its high efficiency, low cost, andease of construction. However, the present invention takes a differentperspective on the relative size of the individual engine types.Although modem gas turbines have efficiency levels in the 30 to 40%range (LHV), they require the use of an HRSG and steam turbine toachieve the combined cycle efficiency of 50 to 60% (LHV). In addition,to effectively recover the heat of the exhaust gases, these HRSGstypically have three pressure levels for the steam, high-pressure,intermediate pressure, and low pressure. The use of the intermediate andlow-pressure steam results in an overall steam cycle efficiency of only34 to 36%.

[0275] Modern large power plant steam cycle efficiencies, however, arein the 45% to 50% efficiency range. To achieve these levels, the use oflow-pressure steam, as is the case with conventional combined cycles, isunacceptable. Therefore, several preferred embodiments of the presentinvention describe a method that utilizes only high-pressure steam toachieve high steam cycle efficiencies in a combined cycle configuration,yet still recovers as much heat from the exhaust gases of the gasturbine as the high efficiency, combined cycle technology in the priorart.

[0276] By this implementation, the new technology combined cycle powerplant diverges from the typical arrangement in the prior art where thegas turbine (GT) was the prime (larger) engine and the steam turbine togas turbine power ratio was approximately 1:2, to an arrangement wherethe steam turbine (ST) is typically the prime (larger) engine and the STto GT power ratio (ST/GT) can typically be selected to be in the rangeof 0.75:1 to 2.25:1 or greater. This ratio is easily adjusted by thedesign of the steam turbine, the rated amount of supplemental firing,and the steam cycle.

[0277] During the operation of any power plant, the operations staffmust modulate the power plant's output to the load on the system (powerconsumption by all users in the electrical grid). As the system loadfluctuates, the total power produced by all the power plants connectedto the grid must change to meet this fluctuation, otherwise, the speedof the equipment will change, and the resulting power produced will nolonger be at 60 Hz (60 cycles per second for U.S. plants, etc.). Thiswill have a dramatic effect on the equipment that the end users have inservice (i.e. electric clocks will not keep accurate time, electricmotors will not operate at appropriate speeds, etc.). Therefore, theutility and power plant personnel are responsible for maintaining aconstant frequency or speed on their equipment. To achieve this, theymust constantly change their power output to match that of the system.Note in European and various other countries this standard frequency is50 Hz, versus 60 Hz in the United States and other countries in theWestern Hemisphere such as Canada.

[0278] During the hot summer months and on extremely cold days in thewinter, the system load is near its seasonal peak. Also, typicallybetween 4 PM and 8 PM on weekdays, the system is near its daily peak.However, during nights and weekends, the system load might only average60% of the weekday peak. Due to these dynamics for the system load, itis uncommon for a dispatched power plant (dispatched means controlled bythe utility to meet system load) to operate at its rated output, or anysteady load, for an extended period of time. Instead, it is typicallyoperated at high loads during weekday peak hours (not necessarily itsrated output) and at relatively low loads (approximately 60% output) forextended hours during nights and weekends. Refer to FIGS. 31A, 31B, and32 for more information on typical load profiles.

[0279] Therefore, to be efficient, a power plant must have theflexibility to operate continuously at varying loads between 50% and100%. Conventional combined cycle power plants are efficient, butsacrifice a great deal of efficiency when operating at part load. Thisis especially true of plants where the GT is the primary engine. Inthese plants, to reduce load initially from full load, the moresophisticated GTs equipped with inlet guide vanes (IGVs) will reduceairflow through the engine, thus reducing their pressure ratio. Inaddition, to further reduce load, these engines must reduce theirturbine inlet temperatures (also referred to as firing temperature) tooperate at part load. Reducing these pressures and temperatures greatlyreduces the operating efficiency of the GT engine.

[0280] To improve combined cycle plant efficiency, reduce cost, loweremission levels, reduce the plant real estate requirements, and simplifyoperations and maintenance (O&M), the present invention teaches the useof an HRSG optimized for continuous supplemental firing that utilizes asingle pressure level evaporator (boiler) with equal or greater ST inletpressures than are typically employed in combined cycle applications inthe prior art. In addition, it proposes the use of some features used inconventional Rankine Cycles not employed in conventional combinedcycles.

[0281] Refer to FIG. 9 for an exemplary embodiment of this new cycle. Asin a typical combined cycle application in the prior art, this newarrangement utilizes one or more GTs (920) as the topping cycle powerdevice. Also, as in the typical combined cycle application in the priorart, the GT exhaust gases are fed into the HRSG. From this point,however, the cycle is changed from conventional combined cycle practice.A single pressure level HRSG is utilized rather than a multiple pressurelevel HRSG. To maximize cycle efficiency, the pressure of steam producedcan be much higher than the nominal 1800 psia typically seen. Thispressure could be supercritical (greater than 3206 psia) if desired. Forsimplicity, this discussion will focus on a sub-critical application(2400 psig rating) with an exemplary implementation example. However,performance curves for supercritical steam conditions will be includedand discussed.

[0282] Energy Flow Analysis

[0283] First, it is instructive to examine the overall energy flow in aconventional combined cycle application. From a simple energy analysis,FIG. 21 illustrates the energy flow in a combined cycle applicationwhile FIG. 10 quantifies, for the Prior Art option, the flow of energyin a conventional combined cycle plant (see the subsequent section onPreferred Embodiment Cycle Optimization for the equations used tocalculate the values in FIG. 10). This table documents performance for aGE model PG7241(FA) GT at ISO conditions with 3.0 inches H₂O inlet airpressure drop and 10.0 inches H₂O exhaust pressure drop. ISO conditionsare defined as 59° F. and 14.696 psia ambient pressure. Referring toFIG. 21, of the initial fuel input to the GT, GTI (2101) 32.31% (allpercentages based on HHV) is converted into electricity, which is the GToutput, GTO (2105). Based upon the GT exhaust gas flow and its enthalpy,only 56.21% of the input energy is sent to the HRSG, HGI (2103), meaningthat 11.48% is lost between the GT and the HRSG GTL (2102). This islikely GT generator losses, GT heat loss, gear driven accessories, motordriven accessories, windage loss, and other miscellaneous losses. Forthis example, no energy from supplemental firing will be added,therefore, SFE (2104) is zero. Of this remaining 56.21% of the GT inputenergy sent to the HRSG HGI (2103), about 10.7% (which equals(0.107)(0.5621) or 6.04% of initial GT input energy) is lost up theexhaust stack HGE (2107).

[0284] Of this remaining energy available in the GT exhaust gases toproduce steam in the HRSG, 1% is considered to be lost as heat toambient HGL (2106). Converted into terms of GT input energy, thisequates to losses of 6.04% of the GT input energy for exhaust loss and0.50% of the GT input energy for HRSG heat loss. This now leaves 49.67%of the GT input energy as energy transferred to the steam HRS (2108)which is available for recovery and conversion to electricity by the ST.

[0285] With a published heat rate of 6040 BTU/kWh (LHV) for a GE STAG™S207FA plant with two GE Frame 7s and one ST, the plant efficiency basedon the higher heating value (HHV) of natural gas is 50.90%. If the GTconverts 32.31% of the fuel input GTI (2101) into electricity, then bysubtraction the ST must convert 18.59% of the fuel input GTI (2101)input energy into electricity. With a steam turbine generator efficiencyof 99% (or 1% loss, SGL (2110)), and a auxiliary load factor of 97.5%,and 49.67% of the fuel input HRS (2108) available to the ST cycle, thenthe basic steam cycle efficiency calculates to 38.78%((18.59/49.67)/(0.975)(0.99)). This is significantly less than the46.78% efficient operation from advanced steam cycles in a Rankine Cycleonly plant (see FIG. 4).

[0286] This steam cycle efficiency is confirmed by General Electric intheir informative document GER-3574F (1996), entitled “GE Combined-CycleProduct Line and Performance” by David L. Chase, Leroy O. Tomlinson,Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak. In discussingsupplemental firing of the HRSG, this document states

[0287] “ . . . the incremental efficiency for power production bysupplemental firing is in the 34-36% range based upon the lower heatingvalue [(LHV)] of the fuel.”

[0288] Since supplemental firing adds heat only to the steam cycle, ittherefore follows that the steam cycle efficiency of GE's combined cycleplants is as stated.

[0289] Cycle Efficiency

[0290] The next question to be answered is how does one achieveconventional steam plant cycle efficiencies with the steam portion of acombined cycle? In a review of FIG. 4, it should be observed that reheathelps improve steam cycle efficiency. However, reheat is alreadyemployed by many of the high efficiency combined cycles, such as the GESTAG™ plant S207FA which utilizes two GE Frame 7s and one ST to achievea heat rate of 6040 BTU/kWh LHV (refer to GE informative documentGER-3574F (1996), entitled “GE Combined-Cycle Product Line andPerformance” by David L. Chase, Leroy O. Tomlinson, Thomas L. Davidson,Raub W. Smith, and Chris E. Maslak). Therefore, other steam cycleefficiency enhancements are the use of higher inlet pressures, a highervolumetric flow steam turbine (for higher ST efficiency) and feedwaterheating. These enhancements will be applied in several preferredembodiments of the present invention.

[0291] Most modem high efficiency GTs such as the GE model PG7241(FA)have internal firing temperatures near 2400° F. These GTs are designedto have exhaust gas temperatures at rated load in the range of 1100° F.For the PG7241(FA), at ISO conditions corrected for an HRSG exhaust lossof 10.0 inches H₂O, the exhaust gas temperature is 1123° F. Thecorresponding exhaust gas flow is 3,552,000 lb/hr. At 1800 psia inletpressure, 1050° F. inlet temperature, with reheat to 1050° F.,exhausting at 1.2 inches of mercury absolute (HgA), this steam cyclewould require 1642.4 BTU/lb of heat input (reference FIG. 4, Option 2).With an inlet enthalpy of the exhaust gases of 412.6 BTU/lb, and exhaustgas enthalpy of 159.2 BTU/lb, the exhaust gases have the energy contentto produce 548,000 lb/hr of steam flow. However, closer examinationreveals a flaw in this logic. At 1800 psia, steam boils at 621° F. Sinceheat flows from higher temperatures to lower temperatures, a reasonabletemperature for the exhaust gases leaving the evaporator section wouldbe 650° F. If water preheated to an enthalpy of 648 BTU/lb was suppliedto the evaporator (an optimistic assumption), the energy required forsteam production would be 1642.4−648=994.4 BTU/lb.

[0292] However, at 650° F., the exhaust gases have an enthalpy of 281.3BTU/lb. Therefore, the exhaust gases have the ability to boil and reheatonly 469,000 lb/hr (3,552,000)(412.6-281.3)/994.4. Hence, the issuebecomes a heat exchange problem, as there is insufficient high level(high temperature) energy to provide the steam at higher pressures.

[0293] Conversely, from 650° F. exhaust gas temperature to exhaust at180° F., there is sufficient energy to preheat 729,000 lb/hr of water((3,552,000)(281.3-159.2)/(648-53)) from the hotwell at 53 BTU/lb tosaturation enthalpy of 648 BTU/lb. Therefore, for heat recovery, in theprior art, the single pressure boiler is inefficient and either makesexcess hot water, which has little or no use in a power productionfacility, or has an HRSG exhaust temperature that greatly exceedsoptimum. This result has prompted the introduction of the multi-pressurelevel HRSG within the prior art. This arrangement makes use of theaforementioned hot water or exhaust gas energy by providing a lowerpressure evaporator section(s) in the HRSG that converts what would benon-usable hot water/exhaust gas energy to lower pressure steam.Although it has less energy content and less ability to produce power inthe ST than the high-pressure (HP) steam, this low-pressure (LP) steamnonetheless does add to the power output of the ST and serves to reducethe plant's heat consumption for a given power output (heat rate).

[0294] Supplemental Firing

[0295] Another method to alleviate the heat transfer shortcomings of asingle pressure HRSG, without adding more pressure levels as in theprior art, is to add energy at or near the inlet of the HRSG throughsupplemental firing. However, the current teachings are thatsupplemental firing reduces overall plant thermal efficiency. This isnoted by Moore in U.S. Pat. No. 5,649,416 in which he states

[0296] “Supplemental firing of the heat recovery steam generator canincrease total power output and the portion of the total power producedby the steam turbine, but only with a reduction in overall plant thermalefficiency.”

[0297] In addition, Rice, in U.S. Pat. No. 5,628,183 states

[0298] “Supplementary firing in front of the HRSG does not offer aviable solution towards higher cycle efficiency.”

[0299] Thus, the prior art specifically teaches away from this techniqueof supplemental firing. In addition, Rice references other documents byGE and Westinghouse that concur with his statement. GE informativeliterature, GER-3574F (1996), entitled “GE Combined-Cycle Product Lineand Performance” by David L. Chase, Leroy O. Tomlinson, Thomas L.Davidson, Raub W. Smith, and Chris E. Maslak states

[0300] “ . . . the incremental efficiency for power production bysupplemental firing is in the 34-36% range based upon the lower heatingvalue of the fuel.”

[0301] This states that although combined cycle efficiency is 56% basedon the lower heating value (LHV) of the fuel at full load, powerproduced through supplemental firing is added at an efficiency equal toor less than 36% LHV.

[0302] Also in this document, (GE informative document GER-3574F, 1996,entitled “GE Combined-Cycle Product Line and Performance” by David L.Chase, Leroy O. Tomlinson, Thomas L. Davidson, Raub W. Smith, and ChrisE. Maslak), another source which identifies supplemental firing as adetriment to efficiency (heat rate) is Table 14 which indicates thatHRSG supplemental firing can increase combined cycle plant output in theprior art by 28%, but only with an increase in overall combined cycleheat rate (specific fuel consumption) of 9%.

[0303] Present Invention Energy Flow

[0304] It will now be instructive to reexamine the overall energy flowin a combined cycle application as utilized in several embodiments ofthe present invention. From a simple energy analysis, FIG. 21illustrates the energy flow in a combined cycle application while FIG.10 quantifies the energy flow in a preferred exemplary embodimentcombined cycle (see the following section on Preferred Embodiment CycleOptimization for the equations used to calculate the values in FIG. 10).Again, GT performance is for a GE model PG7241(FA) GT at ISO conditionsand 3.0 inches H₂O inlet air pressure drop and 10.0 inches H₂O exhaustpressure drop. Referring to FIG. 21, of the initial fuel input to theGT, GTI (2101), 32.31% (all percentages based on HHV) is converted intoelectricity, which is the GT output, GTO (2105). Based upon the GTexhaust flow and enthalpy, only 56.21% of the GT input energy is sent tothe HRSG HGI (2103), meaning that 11.548% is lost between the GT and theHRSG, GTL (2102). Of this remaining 56.21% of the GT input energy sentto the HRSG, about 10.7% of it is lost up the exhaust stack HGE (2107),leaving 50.17% of GT input energy to the HRSG. To this point, the energyflow is unchanged from the prior art.

[0305] To ensure maximum heat recovery in the HRSG, several of thepreferred embodiments of the present invention prescribe increasing thefeedwater flow through the HRSG until there is a sufficient balance ofheat gain by the feedwater to match the necessary heat loss from theexhaust gases for optimum heat recovery (i.e. reduce HRSG exhaust gastemperature to approximately 180° F.). Secondly, through the addition offuel at the HRSG inlet (supplemental firing), the exhaust gas energy inthe HRSG is raised until there is sufficient energy to convert most orall the feedwater flow into steam. Using Option 3 from FIG. 4, heat mustbe added at 1633.9 BTU/lb to produce the desired steam conditions. Sincethe heat capacity of the exhaust gases is approximately 0.25 BTU/lb/°F., and the heat capacity of the returning condensate is approximately1.0, the steam flow should be near 0.25 lb of steam per lb of exhaustgas flow. For two GE frame 7 GTs this yields a steam flow of 1,776,000lb/hr.

[0306] To produce this amount of steam will require 2894 MMBTU/hr(million BTU/hr). With 1% loss to ambient from the HRSG, HGL (2106), theheat input requirement becomes 2923 MMBTU/hr. With exhaust loss, thenecessary HRSG input energy required to produce this steam is 87.99% ofthe GT input energy. Since the HRSG input energy HGI, (2103) minus theHRSG exhaust loss, HGE (2107), equals 50.17% (56.21(−6.04), of GTI(2101), an additional amount of energy equal to 31.78% of the GT inputenergy must be added as heat from supplemental firing SFE (2104),yielding a total of 81.95% of GTI (2101). Adjusting for a 1% loss toambient, HGL (2106), 81.13% of GTI (2101), the GT input energy, convertsto steam. This steam is now available for conversion to electricity bythe ST.

[0307] With a ST for use in a preferred exemplary embodiment, higherpressure, reheat, and feedwater heating may all be employed. Inaddition, the ST rating will be an estimated 2.5 times that of aconventional combined cycle plant. This would lead a reasonable designerto use the steam cycle efficiency of 44.39% as shown for Option 5 inFIG. 4 (as per Moore, ST efficiencies increase with rating, but fordemonstration purposes, an overall 90% has been retained for thisexample).

[0308] Utilizing a 44.39% efficient basic steam cycle, 36.01% of theavailable heat is converted to shaft horsepower, utilizing a factor of97.5% to account for auxiliary loads and a 99% efficient generator. STelectrical output is therefore 34.76% of GTI (2101), GT input energy.With the GT output, GTO (2105), equal to 32.31% of GTI (2101), the SToutput equal to 34.76% of GTI (2101), and with an additionalsupplemental fuel input of 31.18% of GTI (2101), combined cycleefficiency therefore becomes (output divided by input)((0.3231+0.3476)/(1+0.3118)) which equals 50.90%. Utilizing only two (2)FWHs in the cycle, the efficiency of this exemplary preferred embodimentis on par with the GE conventional combined cycle plant. Forsupercritical applications, the overall combined cycle efficiency inseveral of the preferred embodiments increases to 51.75% and lowers theheat rate to 5942 BTU/kWh LHV (reference FIGS. 10 and 21).

[0309] Therefore, from an overall energy perspective, it is apparentthat supplemental firing is NOT detrimental to overall combined cycleefficiency IF a commensurate increase in bottoming cycle efficiencyaccompanies the supplemental energy addition to the bottoming cycle.

[0310] Preferred Embodiment Cycle Optimization

[0311] As stated, one of the major improvements for several of thepreferred embodiments of the present invention is the flexibility. Withsupplemental firing, the new combined cycle power plant can be designedwith a combination of various gas turbines together with a customdesigned steam turbine(s) to provide a much wider range of applicationfor the new combined cycle power plant.

[0312] Since efficiency is defined as output divided by input, theenergy flow analysis can be used to determine the steam cycle efficiencyrequired at a given rating. Therefore, for overall combined cycleefficiency, the output is the combination of both the steam turbine andgas turbine(s) electrical output. The input is the total of the GT inputenergy along with the energy added to the duct burners throughsupplemental firing. Therefore, referring to FIG. 21, the equation forcombined cycle efficiency (.eta.) of several of the preferredembodiments of the present invention is given by the relation:$\begin{matrix}{\eta = \frac{\left( {{GTO} + {STO}} \right)}{\left( {{GTI} + {SFE}} \right)}} & (2)\end{matrix}$

[0313] where

[0314] GTO=gas turbine(s) electrical output

[0315] STO=steam cycle electrical output

[0316] GTI=gas turbine(s) input energy

[0317] SFE=HRSG input energy through supplemental firing.

[0318] In the above exemplary equation, the values of GTO, GTI, and SFEare typically known. The unknown variable is the steam cycle electricaloutput, STO. This number is a function of several other inputs,including steam turbine generator efficiency, HRSG exhaust loss,auxiliary load factor, and finally steam cycle efficiency. First, it isnecessary to calculate the amount of energy that is transferred to thesteam from the HRSG. This is defined as HRS (2108) and is calculatedfrom the following equation:

HRS=[HGI+SFE−HGE] (1−HGL)  (3)

[0319] where

[0320] HRS=HRSG energy transferred to steam

[0321] HGI=GT exhaust heat

[0322] SFE=supplemental firing heat

[0323] HGE=HRSG exhaust loss

[0324] HGL=heat loss to ambient

[0325] The above exemplary equation essentially calculates the heat intothe steam as the sum of: the GT exhaust heat, plus the heat added fromsupplemental firing, minus the HRSG exhaust loss, with a correction forheat loss to ambient. This now defines the quantity of energy availableto the steam cycle. To determine the electrical output from this energy,STO (2111), this energy input must be adjusted for the steam cycleefficiency, SCE (2109), the ST generator losses, SGL (2110), and theauxiliary loads, AXF (2112). The equation for steam turbine generatoroutput becomes:

STO=HRS×SCE×AXF×(1−SGL)

[0326] where

[0327] HRS=HRSG energy transferred to steam

[0328] SCE=steam cycle basic efficiency

[0329] AXF=auxiliary load factor

[0330] SGL=steam turbine generator losses=(1−steam generator efficiency(SGE))

[0331] The steam cycle efficiency value therefore converts steam energyinto ST shaft power, which is then corrected to steam cycle electricaloutput by corrections for both the generator efficiency and thereduction of power output by the auxiliary loads.

[0332] Knowing these equations, and also knowing the desired output fora given GT arrangement (see FIG. 29 for range of outputs of several ofthe preferred embodiments of the present invention), the required steamcycle efficiency can be determined which will yield a preferredembodiment combined cycle plant efficiency equal to that of theconventional (lower rating) combined cycle plant from the prior artwhich was based on the same GTs. FIG. 30 illustrates the steam cycleefficiencies that are required as the combined cycle plant described byseveral of the preferred embodiments of the present invention isincreased in rating. Note that the parameter along the horizontal axisis the ratio of ST power output to the total of all GT(s) power output.

[0333] Utilizing FIG. 29, FIG. 30, and the aforementioned equations forsteam cycle efficiency and overall plant efficiency, a design engineerskilled in the art can determine which GT combination is most favorablefrom both an energy efficiency and economic perspective, and determinethe relative complexity of the steam cycle (subcritical steamconditions, amount of feedwater heating, inlet temperatures, etc.) thatwill yield the desired overall plant efficiency. Refer to FIGS. 47-50illustrating a plant design/construction method for the selection,design, and financing of the preferred embodiment of the presentinvention.

[0334] Preferred Embodiment Flexibility

[0335] As previously mentioned, flexibility is one of the majoradvantages to the present invention. From an examination of FIG. 30, itcan be seen that at lower ST/GT ratios, a steam cycle of more moderateefficiency can be utilized to provide on par plant efficiency utilizingthe teachings of the present invention. However, it would be possible atlow ST/GT ratios to utilize ultrasupercritical steam conditions toexceed the efficiency of a combined cycle power plant from the priorart. If the exemplary preferred embodiment in FIG. 26 at 725 MW were touse an ultrasupercritical bottoming cycle, the heat rate would bereduced from 6006 BTU/kWh to 5912 BTU/kWh.

[0336] However, unlike preferred embodiments with higher ST/GT ratios,this configuration yields less operational flexibility than preferredembodiments with higher ST/GT ratios. With these lower ratios, thecontrol of the preferred embodiment will be more like that of the priorart in that the GTs will need to be modulated to control plant load at ahigher plant operating point. Depending upon the economics, highefficiency, low efficiency, or capital costs will determine which ST/GTratio is ultimately chosen by the power plant developer.

[0337] Preferred Embodiment Potential Ratings and ST-GT Ratio

[0338]FIG. 30 illustrates the approximate steam turbine rating increasesthat are attainable from several of the preferred embodiments of thepresent invention. A conventional combined cycle power plant from theprior art could have a ST output that is nominally 55% of the total GTsoutput. Therefore, total plant output could be defined as 1.55 (1.00 forGTs plus 0.55 for the ST) of GTs output. With this example of several ofthe preferred embodiments of the present invention, the ST could bedesigned to be as much as 2.1 times the output of the GTs, such thattotal plant output is 3.1 (1.0 for GTs plus 2.1 for the ST) times theoutput of the GTs.

[0339] This example of a preferred embodiment of the present inventionhas a rating that is 3.1/1.55=2.0 times that of the prior art. To remainon par in efficiency with the prior art, however, the basic steam cycleefficiency needs to be 48.3% (refer to FIG. 30). With supercriticalsteam conditions, advanced steam parameters, and feedwater heating,basic steam cycle efficiencies can come close to this benchmark.Therefore, several of the preferred embodiments of the present inventionhave the ability with certain gas turbine arrangements to nearly doublecombined cycle power plant output as compared to the prior art,drastically reduce the amount of hardware that would have been requiredin the prior art to attain this output, yet still manage to achieveefficiency levels that are on par with the prior art.

[0340] Since the present invention teaches the use of a single pressurelevel HRSG, and to efficiently utilize a single pressure level HRSG, thefeedwater flow through the low temperature section of the HRSG must beadequate to absorb the GT exhaust gas energy, analysis has shown that anST/GT ratio minimum of approximately 0.75 is required to meet thisobjective. Assuming a relative GT power output of 1.0 and a ST/GT powerratio of 0.75, yields a total plant power output of 1.75 times the GToutput, resulting in a GT to total power output of (1.0/1.75) orapproximately 0.57 or 57% of the total plant power output.

[0341] Design Limitations

[0342] Although several of the preferred embodiments of the presentinvention offers a more expansive range of combined cycle ratings for agiven set of gas turbines than was available in the prior art, there arestill limitations on the design of these new technology combined cyclepower plants. Some of these limitations are as follows:

[0343] 1. Above approximately a 1600° F. duct-fired gas temperature atits inlet, the HRSG will require a more expensive water-wallconstruction.

[0344] 2. With water-wall construction, the HRSG may be limited toapproximately a 2400° F. duct-fired gas temperature.

[0345] 3. The HRSG exhaust gases must contain sufficient oxygen tosupport the combustion of additional fuel.

[0346] 4. The duct burners that provide additional heat input to theHRSG must be able to maintain low NOX levels even at high prescribedfiring rates.

[0347] 5. The cycle must be designed to operate within the steam turbinedesign parameters for pressure and temperature.

[0348] 6. The cycle must be designed so as to maintain the properefficiency, cost, emissions, or other limiting parameters that may existto make the project economically and environmentally acceptable.

[0349] Considering these limitations, FIG. 29 illustrates an approximaterange of rated power for combined cycle power plants described byseveral of the preferred embodiments of the present invention. Note thatthese power plants are based upon either one or two GTs and cover arange from less than 150 MW up to 1050 MW. FIG. 29 is not meant torepresent all possible GT combinations which can utilize the preferredembodiment of the present invention, but represents only a sample ofvarious GTs for demonstration purposes.

[0350] Impact of Economic Considerations on Plant Design

[0351] All power plant design engineers skilled in the art reviewnumerous plant design options for their relative economic merit beforeselecting a final configuration for a power plant. This is true with thecombined cycle plants from the prior art and will be true of combinedcycle plants utilizing the system and method described by several of thepreferred embodiments of the present invention. New plants must becommercially feasible if they are to be constructed.

[0352] The power plant design engineer may examine alternatives such asa low cost cell type cooling tower with a high auxiliary load (electricmotor driven fans) versus a high cost hyperbolic style cooling towerwith only a small auxiliary load (natural draft air flow, no fansrequired). This becomes an economic evaluation of the energy savedversus the capital cost expended to save said energy. Based upon currentand projected economic factors for energy costs, capital costs, andother factors, the developer of the power plant project will select themost economical arrangement. The most efficient selection from an energyconservation perspective is not always the most economical selection.

[0353] These same type of evaluations will need to be presented withseveral of the preferred embodiments of the present invention. Althoughultrasupercritical steam conditions yield higher steam cycleefficiencies, the incremental savings in fuel may not outweigh the addedcost for the more intricate hardware. If interest rates are high,several of the preferred embodiments of the present invention will allowlarge capacity increases with only a nominal percentage increase inprice. With low fuel costs, larger plants without the commensurateincrease in steam cycle efficiency may be appropriate. Again, several ofthe preferred embodiments of the present invention allow the designengineer skilled in the art along with the plant developer to chose froma wider range of alternatives to find the most commercially viableoption for the power plant.

[0354] With several of the preferred embodiments of the presentinvention becoming primarily a steam plant rather than primarily a GTplant, there are a couple economic evaluations that are usually of keyinterest. Since these steam turbines will be large and have high exhaustend flows, they typically utilize either one, two, or three exhaustcasings, each of which has a double flow arrangement. FIG. 51 is anillustration of a General Electric (GE) steam turbine a from GEinformative document entitled “Steam Turbines for Large PowerApplications” by John K. Reinker and Paul B. Mason (General ElectricReference GER-3646D, 1996). The casing to the left is the combined HP/IPsection, while the two larger sections to the right are the double flowexhaust (LP) sections. Differing sizes of exhaust casings are availablewhich are designed around the blade lengths in the last stage. There canbe substantial cost differentials between the different exhaust casings.

[0355] The selection of the steam turbine last stage blade height,exhaust casing size, and number of exhaust casings is one very commoneconomic evaluation for a large steam plant. The steam cycle may becomemore efficient by an increase to the next larger exhaust casing orperhaps even through the addition of another exhaust casing. However,the incremental increase in steam cycle efficiency must be weighedagainst the increase in cost for the additional hardware. Another factorthat comes into play is the sizing of the condenser and heat rejectionequipment. Again, lower exhaust pressures yield higher steam cycleefficiencies, but the cost of the equipment to provide incrementalreductions in exhaust pressure must not outweigh the fuel savings.

[0356] In consideration of the economics of operation, the developermust provide the design engineer with an operation scenario for the newpower plant. Since the system electrical load is very dynamic andconstantly changing, a load profile needs to be established whichexemplifies the load on the plant as a function of time. FIG. 31A isfrom a U.S. Department of Energy report numbered DOE/EIA-0614 entitled“Electricity Prices in a Competitive Environment: Marginal Cost Pricingof Generation Services and Financial Status of Electric Utilities”. FIG.31A illustrates a typical load profile for a system (electrical grid) onan hourly basis for a single day. On a weekly basis, this profile wouldindicate lower load on weekends and holidays, and on an annual basis,there would be adjustments for seasonal changes. Since most power plantswill operate a majority of their lifetime at partial load, the optimumeconomical arrangement results from designing these plants to be mostefficient at some average or mean load point of operation, versus at theplant's rating.

[0357] This is noted by M. Boss in GE informative document GER-3582E(1996), entitled “Steam Turbines for STAG™ Combined-Cycle PowerSystems”. In this paper, the author explains that although theefficiency of the steam cycle may be maximized when the ST exhaustannulus velocity at the last stage blade is approximately 550 feet persecond, the economic optimum is typically with an exhaust annulusvelocity of between 700 and 1000 feet per second at the rated point ofthe ST. James S. Wright, in GE informative document GER-3642E (1996),entitled “Steam Turbine Cycle Optimization, Evaluation, and PerformanceTesting Considerations” provides an evaluation for steam turbine exhaustcasing selection. In the example, the selection is made between threedifferent sized exhaust casings, with the efficiency of the exhaustcasings increasing with each larger size. The largest casing is notselected because its incremental gain in efficiency does justify itsadded cost per the economic parameters. By the same token, the smallestcasing is not selected because its incremental savings in capital costdoes not justify the large loss in efficiency. Therefore, the mediumsized casing is selected because it is the economic optimum.

[0358] Single Pressure HRSG

[0359] To make an HRSG effective at a single pressure level, its designeffectiveness must first be examined. FIG. 11 is a curve of steamenthalpy versus temperature for a pressure of 1800 psia. As can bereadily seen, the heat content of the steam is not a linear functionwith respect to temperature. This phenomenon greatly complicates theheat transfer with the exhaust gases that have a nearly linearcharacteristic (see FIG. 12). As is seen in FIG. 11, at the boilingpoint of 621° F., the water/steam mixture increases in enthalpy from 648BTU/lb to 1154 BTU/lb without any increase in temperature. The heatabsorbed in this section of the HRSG (evaporator) will be much greaterthan any other section for a given temperature change.

[0360] Between the temperatures of 100° F. and 400° F., the average heatcapacity of water is 1.014 BTU/lb/° F. This value is essentially linearand changes only slightly with pressure. Therefore, heat transfer inthis region between the water and exhaust gases will be relativelyconsistent.

[0361] To maximize the effectiveness of heat recovery in the HRSG, andto still provide the maximum amount of steam to the ST, a system controlmethod is required that optimizes the feedwater/steam flow through eachsection of the HRSG. This optimization scheme will be programmed in thepower plant's DCS control system.

[0362] System Control

[0363] There are numerous possible control techniques for the ST,however, two popular methods are flow control and sliding pressurecontrol. With flow control, the ST includes a set of valves that iscontrolled to maintain design inlet pressure. With sliding pressureoperation, the inlet pressure to the ST is allowed to “slide” or changewith the load (steam flow) to the ST. For combined cycle plants, whereheat recovery is employed, it is often advantageous to use slidingpressure control. This control method allows for high volumetric flowsin the steam turbine by utilizing lower specific volume steam (lowerpressure) at part load. This maintains ST efficiencies at or near designlevels. In addition, lower pressure steam boils at a lower temperaturethan higher pressure steam, therefore, the lower temperature exhaustgases in the HRSG associated with lower loads can produce more steam.

[0364] Energy Utilization

[0365] As demonstrated previously, in order to produce high-pressuresteam in the HRSG, it is necessary not only to have the overall energycontent to produce the steam (total required BTUs), but the energy mustbe at the appropriate temperature to affect the necessary heat transfer.In addition, it is desired to maximize the use of waste heat, and notproduce large quantities of hot water or greatly increase HRSG exhausttemperature above its optimum point. The use of supplemental firingbecomes extremely useful in meeting these goals.

[0366] It is now useful to consider the concept of a single pressureHRSG used with a GT. As previously demonstrated, this arrangement, whendesigned for an HRSG exhaust temperature of 180° F., would produceeither an excess of hot water or a high HRSG exhaust gas temperature inthe prior art. Again, this is due to the non-availability of sufficientheat input at the higher temperatures, and an overabundance of heatavailable at the lower temperatures. For illustration purposes, consideran HRSG that added heat to the exhaust gases without an increase intemperature (not a likely arrangement for a GT/HRSG assembly). Imaginethat to add heat, the HRSG was designed to ingest more fuel and moreair, but without an increase in its inlet temperature. This scenariowould provide for the production of a larger quantity of GT exhaustgases and thus a larger quantity of steam. However, it also wouldprovide for a larger quantity of hot water. Effectiveness of the HRSGwould not be changed, only its capacity would be increased proportionalto the heat addition.

[0367] This concept is important, because not only does it apply to theHRSG, but it applies to the conventional combined cycle practice whensupplemental firing is utilized. In the prior art, supplemental firingincreased steam flows, but did not improve the effectiveness of thesteam cycle.

[0368] Due to excess oxygen in the exhaust gases from a GT (oxygenlevels reduce from 21% O₂ in ambient air to approximately 12-15% at theGT exhaust at full load), fuel can be burned directly in the HRSGwithout the need for additional air. This practice allows supplementalfiring to increase the temperature of the exhaust gases. The combustionand heat recovery process for supplemental firing is essentially 99%efficient, as only 1% of the HRSG heat input is lost to ambientsurroundings. This is a dramatic improvement over conventional Rankinecycle boilers that might only be 80 to 90% efficient. The primary reasonfor this large differential in efficiency between the conventionalRankine cycle boilers and HRSGs is that conventional Rankine cycleboilers ingest cold ambient air for combustion and may then exhaust inthe range of 350° F. to 400° F., versus the HRSG which receivespreheated GT exhaust gases at temperatures between 800° F. and 1200° F.,and then exhaust in the range of 160° F. to 200° F.

[0369] This increase in the energy level of the exhaust gases as aresult of supplemental firing, greatly improves the ability (heattransfer capability) of these exhaust gases to produce high-pressure,high temperature steam. In addition, more energy at the high end of theHRSG offsets or balances the excess energy at the low end of the HRSGtypical in the combined cycle from the prior art.

[0370] In other words, additional heat input at the HRSG inlet thatincreases the exhaust gas temperature, can be transferred to thefeedwater flow that had insufficient energy to become HP steam. Not onlyis the overall steam flow increased, but the effectiveness of the steamcycle is also increased by producing a higher proportion of HP steam.Thus, the addition of fuel into the bottoming cycle, as well asproviding additional heat input, can be used to increase the overalleffectiveness of the bottoming cycle.

[0371] System Overview

[0372]FIG. 13 is a conceptual schematic for a combined cycle applicationwith heat addition to the bottoming cycle. In FIG. 13, the topping cyclefluid (TCF) (1301) enters the topping cycle engine, (TCE) (1302) wherefuel and/or heat (CFT) (1303) is added to raise its temperature. Thefluid performs work that is converted by the topping cycle engine intoshaft horsepower. This shaft horsepower drives the topping cycle load,(TCL) (1304). This load could be an electrical generator, pump,compressor, or other device that requires shaft horsepower. Theexhausted fluid from the topping cycle engine is directed through andexhaust line (1305) to a heat recovery device (HRD) (1306). In addition,fuel and/or heat (CFB) (1314) is added to the topping cycle fluid at thepoint where it enters the heat recovery device. After passing throughthe HRD, the topping cycle fluid exhausts to an open reservoir (1307).

[0373] For this example, the topping cycle is an open cycle. In otherwords, the topping cycle fluid is taken from a large reservoir anddischarges to that same reservoir. The heat recovery device (1306)captures a portion of the topping cycle exhaust energy and transfers itto the bottoming cycle fluid (BCF) (1308). In this example, thebottoming cycle fluid is heated at a single pressure level, ahigh-pressure (HP) line (1309). This fluid then travels to the bottomingcycle engine (BCE) (1310) where it produces shaft horsepower to drivethe bottoming cycle load (BCL) (1311). Again, this load could be anelectrical generator, pump, compressor, or other device that requiresshaft horsepower.

[0374] From the bottoming cycle engine, the bottoming cycle fluid entersa heat exchanger (HEX) (1312) where heat is rejected. The bottomingcycle fluid then enters a fluid transfer device (FTD) (1313) where it isthen returned to the heat recovery device. For this example, thebottoming cycle is a closed cycle, meaning that the bottoming cyclefluid is continuously circulated within a closed loop.

[0375] The present invention exemplary embodiment illustrated in FIG. 13contrasts to FIG. 5 in two major ways:

[0376] 1. Fuel and/or heat is added (1314) to the heat recovery device(1306) which is not added in FIG. 5; and

[0377] 2. In FIG. 13 there is only one fluid, HP fluid (1309), which issupplied to the bottoming cycle engine (1312), versus HP fluid (509), IPfluid (510), and LP fluid (511) in FIG. 5 which are supplied to thebottoming cycle engine (512).

[0378] By utilizing fuel and/or heat addition to the bottoming cycle,not only has the energy to the bottoming cycle increased, but so has thecycle's effectiveness, as now all the IP and LP fluid has been upgradedto HP fluid. This HP fluid has the ability to do more work per unit massflow than either the IP or LP fluids.

[0379] There are a number of different fluids that could be applied tothe conceptual combined cycle arrangement, including water, air, steam,ammonia, refrigerants, mixtures, and many others. The intent of apreferred exemplary embodiment is not to limit the number of cycles usedin the combined cycle, limit the fluids in the combined cycle to anyspecific fluid, limit the fluid pressures that may be utilized, or limitany cycle to being either an open or closed cycle, but to demonstratethat the process of upgrading thermal efficiencies of combined cyclescan often be accomplished through the strategic use of additional fueland/or heat input.

[0380] Heat Transfer Analysis

[0381] HRSG LP Economizer Section

[0382] As mentioned previously, the problem in producing HP steam inconventional combined cycle power plants is the distribution of theenergy between the exhaust gases and the steam being produced. Inaddition, to optimize heat recovery, it is desired to have the exhaustgas temperature at the HRSG exit to be optimum. Therefore, a morein-depth look at the heat recovery process must be made.

[0383] To optimize heat recovery in the lower temperature regions of theHRSG (approximately 470° F. exhaust gas entering temperature to the 180°F. exhaust gas exit temperature range), a sufficient amount of heat mustbe removed by the pressurized feedwater. The average heat capacity ofthe exhaust gases in this range (470° F. to 180° F.) is 0.257 BTU/lb/°F. (note that this value can vary slightly with exhaust gas oxygencontent/amount of supplemental firing). Between the temperatures of 100°F. and 400° F., the average heat capacity of water is 1.014 BTU/lb/° F.Therefore, to obtain an increase in feedwater temperature to correspondto a commensurate decrease in exhaust gas temperature, the flow ratioshould be (1.014/0.257) or 3.95 lbs of exhaust gas per lb of feedwaterin this temperature range of the HRSG. A flow ratio at or near thisnumber will optimize heat recovery for this section of the HRSG. Changesin parameters such as exhaust gas oxygen content, inlet watertemperature, and other factors can be monitored in the plant DCS controlsystem and the optimum feedwater flow through each section of the HRSGcan be calculated and controlled.

[0384] Experience has determined that providing cold water temperaturesat the inlet to the LP economizer section (feedwater directly from thecondenser) can have detrimental effects on the life of the economizercomponents. This is due to corrosion problems in the economizer as aresult of tubes and fins in the economizer being colder than the dewpoint of the exhaust gases of the HRSG. Since these components aretypically constructed of a carbon steel or low alloy steel, thecondensed moisture on the tube and fin surfaces corrodes away thesecomponents and reduces heat exchanger effectiveness. Two common methodsare utilized to alleviate this problem. One is to use a feedwaterpreheater to introduce warmer water into the economizer. The othermethod is to construct a portion of the LP economizer from non-corrodingmaterial, such as stainless steel. Either method is acceptable, and theone selected is usually the one that is determined to be economicallyoptimum.

[0385] HRSG HP Economizer Section

[0386] The HP economizer section of the HRSG heats the feedwater fromapproximately 400° F. (exit of the LP economizer), ideally to thesaturation temperature of the pressure in the evaporator section. Usingan average pressure for this example of 1800 psia, the saturationtemperature at this point is 621° F. In this range, the average heatcapacity of the feedwater is 1.230 BTU/lb/° F. To heat this water, GTexhaust gases will need to enter the section approximately 50° F. abovethe feedwater exit temperature, or 671° F. The average heat capacity ofthe exhaust gases in this range (671° F. to 470° F.) is 0.264 BTU/lb/°F. (note that this value can vary slightly with exhaust gas oxygencontent/amount of supplemental firing). Therefore, for this section ofthe HRSG, the flow ratio should be (1.230/0.264) or 4.66 lbs of exhaustgas per lb of feedwater. Since this flow ratio does not match with theLP economizer optimum flow ratio, an adjustment will need to be made tocompensate for this mismatch (differing optimum flows through eachsection).

[0387] HRSG Evaporator Section

[0388] The evaporator section (sub-critical applications) is unique fromother sections in the HRSG in that its inlet and outlet temperatures areessentially constant (for constant pressure operation). This addsstability to the heat exchange process, and the Log Mean TemperatureDifference (LMTD) fluctuates less with variations in flow than that ofother sections since the outlet temperature is essentially constant. TheLMTD is a non-linear heat transfer variable that is used to determinethe heat transfer capability of a heat exchanger.

[0389] Due to this constant temperature factor, the sections downstreamof the evaporator, the HP and LP economizers, see relatively constant(slight variation with pressure/load) input temperatures. However, inseveral preferred embodiments, supplemental firing will greatly alterthe inlet temperatures to the evaporator section, as well as thesuperheater and reheater sections. These increasing and decreasingtemperatures will determine the steam flow through the HRSG, andultimately, the ST output. Therefore, unlike the economizer sections, anoptimized flow ratio is not truly applicable for the upstream sectionsof the HRSG.

[0390] Since the evaporator section of the HRSG absorbs a major share ofthe heat available, and actually produces the steam, its output ismodulated mostly by the section exhaust gas inlet temperature, which islargely a function of the HRSG exhaust gas inlet temperature. Therefore,the control of this section is done primarily through fuel input.

[0391] HRSG Superheater and Reheater Sections

[0392] These sections are similar in that they both heat steam to ahigher temperature. The superheater section receives saturated steamfrom the evaporator section and heats it to the HP turbine inlettemperature. A desuperheater is used at the exit of this section tocontrol the temperature to the desired value.

[0393] The reheater section receives steam from the HP turbine sectionexhaust and reheats it back to the IP turbine inlet temperature. Adesuperheater can be used at the exit of this section to control thereheat temperature, but does so at a cost in cycle efficiency. This isnoted by Eugene A. Avallone and Theodore Baumeister III in MARKS'STANDARD HANDBOOK FOR MECHANICAL ENGINEERS (NINTH EDITION) (ISBN0-07-004127-X, 1987) in Section 9-24 through 9-25 which states:

[0394] “The attemperation of superheated steam by direct-contact waterspray . . . results in an equivalent increase in high-pressure steamgeneration without thermal loss . . . Usually, spray attemperators arenot used for the control of reheat-steam temperature since their usereduces the overall thermal-cycle efficiency. They are, however, ofteninstalled for the emergency control of reheat steam temperatures.”

[0395]FIG. 14 is a set of curves illustrating the heat requirements forthe superheater and reheater sections as a function of flow. Thesecurves do not include small effects for desuperheating, extractionflows, heat loss in the pipe, or other minor adjustments. Notice thatboth these sections require proportional amounts of heat with flow (STload) changes. Therefore, it may be advantageous, although notnecessary, to build these two sections as one in the HRSG, each with itsown appropriate heat exchange surface area.

[0396] HRSG Surface Areas

[0397] In order to obtain the necessary heat transfer from the GTexhaust gases to the water/steam, it is required that sufficient amountsof heat exchange surface area be provided in each section. Thecontrolling equation that describes this overall heat exchange is

Q=U×A×LMTD

[0398] where

[0399] Q=heat transferred in BTU/hr

[0400] U=overall heat transfer coefficient in BTU/hr/ft²/° F.

[0401] A=total surface area in ft²

[0402] LMTD=log mean temperature difference

[0403] with the log mean temperature difference (LMTD) being defined as

LMTD=(GTTD−LTTD)/ln(GTTD/LTTD)

[0404] where

[0405] GTTD=greater terminal temperature difference

[0406] LTTD=lesser terminal temperature difference

[0407] The terminal temperature differences are

[0408] 1. the temperature of the exhaust gas into an HRSG section minusthe water or steam temperature out, and

[0409] 2. the temperature of the exhaust gas exiting an HRSG sectionminus the water or steam temperature in.

[0410] Obviously, the larger value is the GTTD and the smaller value isthe LTTD. If they are equal, then either one equals the LMTD. If eitherthe GTTD or the LTTD become too small, the surface area, A, must becomevery large to compensate. Since the surface area is essentially thetotal effective surface area of all the tubes and fins in the HRSGsection, adding area adds size, weight, and cost to the HRSG.

[0411] The other factor in the heat exchange equation, U, is based uponthe surface coefficient of heat transfer between the water/steam and thetube inner wall, the heat conductance of the tube material and itsthickness, and the surface coefficient of heat transfer between theexhaust gases and the tube outer wall.

[0412] For general purposes, the controlling factor in this equation isthe surface coefficient between the exhaust gases and the tube outerwall. This is because it is the largest resistance to heat transfer, andlike a group of resistors in series in an electrical circuit, thelargest resistance controls the flow. Therefore, factors that have thegreatest effect in changing the outer heat transfer coefficient are ofthe most concern to engineers designing the HRSG and selecting the areasfor each section.

[0413] From a control standpoint, selection of the areas in each sectionis critical, because once the HRSG is built, these areas cannot bechanged, but become a fixed value. Factors which affect changes in thevalue of U are those which change the velocity of the exhaust gases overthe tube surfaces. The predominant deviation is a change in the exhaustgas flow. Since the GT is a constant volume machine, this occurs withchanges in the ambient air temperature. In addition, it occurs with loadchanges on the GT. If these factors can be minimized, the HRSG can bemore readily designed for operation within a narrow band and betteroptimized.

[0414] As will be illustrated in the example of a preferred exemplaryembodiment of the present invention, the disclosed system and methodallows for GT operation at full load (temperature control) over a widerange of total combined cycle plant load. This contrasts sharply to theprior art that utilized changes in GTs load to modulate the overallcombined cycle plant load. Therefore, at most operating points, the onlysignificant changes in HRSG flow will be attributed to ambienttemperature changes (fuel from supplemental firing adds less than 1% tothe exhaust gas flow). With an ambient temperature range of −20° F. to100° F., the exhaust gas flow would vary approximately 13%. For GTs inthe prior art, load changes alone could account for large changes inexhaust gas flow. The GE Model PG7241(FA) gas turbine, at 55% load,produces only 70% of full load exhaust flow. With ambient changes, thistotal flow change could be only 61% HRSG design flow. This off designflow results in inefficiency in the HRSG and requires design compromisesto accommodate such a wide range of operating conditions.

[0415] Due to the large temperatures in the HRSG as a result ofcontinuous supplemental firing, the LMTDs seen in several preferredexemplary embodiments are greater than those in the prior art,Therefore, the required surface areas are reduced and the overall sizeof the HRSGs may be smaller. This results in a substantial cost savingsin terms of both construction and floor space costs.

[0416] HRSG Controls

[0417] In the prior art, HRSG controls for balancing the heat transferwere limited. Desuperheating controls in the superheater and thereheater were common. Supplemental firing to control the steamproduction is not typically used due to its negative impact onefficiency, and its added cost. Bypasses around some economizer andfeedwater sections were sometimes utilized in the prior art.

[0418] With several of the preferred exemplary embodiments, steamproduction is essentially controlled by the supplemental firing rate.More energy input means more steam output. Multiple duct burner rows canbe utilized for improved section heat transfer control. Multiple ductburner rows allow fuel (heat) input at more than one position along theexhaust gas stream of the HRSG, and with limited heating and subsequentsection cooling at several locations along the HRSG, serves to loweroverall HRSG temperatures (possibly avoiding the more expensivewater-wall construction).

[0419] As with the prior art, desuperheating controls will be used inthe superheater, while desuperheating in the reheater should be limitedto emergency control of reheat steam temperatures. Reheat steamtemperatures can be maintained by careful selection of the HRSG heatexchange areas and by adjustment of trim flow in a split superheaterarrangement. Feedwater flow through the HP economizer is controlled tothe optimum exhaust gas/feedwater flow ratio as is the LP economizerflow. With only one pressure level and six sections, the HRSG in thisexemplary embodiment is much simpler to control and adjust than the12-section, three pressure level boiler from the prior art illustratedin FIG. 6.

[0420] HRSG Comparison—Preferred Embodiment to Prior Art

[0421] The HRSG in several preferred exemplary embodiments may in manycircumstances be similar to the HRSG in the prior art in that it willhave a large number of tubes that transport the feedwater and recoverheat from the GT exhaust gases and transfer it to the water/steam in thetubes through convective heat transfer. This device will be very large.Both a preferred exemplary embodiment and the prior art HRSGs will becontained in a large housing that directs the GT exhaust gases from theGT exhaust to the HRSG exhaust stack. The HRSG may be oriented in eithera horizontal or vertical orientation as required to meet mechanicalconstruction constraints.

[0422] Several preferred embodiments of the present invention, however,will have only one pressure level. This does not exclude the use ofadditional pressure levels, only that single pressure level is exemplaryof a preferred best mode exemplary embodiment. This arrangementcontrasts with the prior art which utilized multi-pressure level HRSGsto maximize heat recovery.

[0423] With only one pressure level and design for continuoussupplemental firing, a preferred exemplary HRSG embodiment may requireless heat exchange area than the prior art. This will serve to reduceoverall size, footprint, weight and cost. Some of the cost savings,however, will be offset by the need for higher temperature materialsand/or perhaps water-wall construction in a preferred exemplary HRSGembodiment.

[0424] With less surface area in a preferred exemplary HRSG embodiment,it is likely that the exhaust backpressure experienced by the GT due tothe HRSG will be reduced. This will serve to increase the GT output andefficiency. Supplemental firing, however, tends to increase thisbackpressure and will reduce some of the performance gains achieved as aresult of lower exhaust gas restriction.

[0425] Due to the flexibility added by a preferred exemplary embodimentto the steam cycle, the GTs will operate at full load over a wider rangeof total combined cycle plant output. This factor serves to provide amore constant flow to the HRSG, provide for a more optimized design, andeliminate inefficient operation at part load conditions.

[0426] With only one pressure level, the HRSG from a preferred exemplaryembodiment will be easier to monitor and control. With only smallchanges in flow and/or temperatures in the HRSG, a preferred exemplaryembodiment is able to make small adjustments in the sectionfeedwater/steam flows to compensate for these changes. With the addedsections, greater variations in exhaust gas flow, and its lesscomprehensive control system, the HRSG in the prior art was more of areactive system to the ever changing system parameters, versus apreferred exemplary embodiment which is more of a proactive system.

[0427] New Overall Combined Cycle Power Plant

[0428] The new overall combined cycle power plant of a preferredexemplary embodiment will be similar to the prior art, but will haveboth subtle and major differences. The major pieces of equipment, theiroperation, and cost impact will now be examined and compared relative tothe prior art.

[0429] Gas Turbines

[0430] The GTs utilized in several preferred exemplary embodiments maybe standard GTs as would be used in the prior art. The only differencewould be from a performance standpoint regarding the amount of pressuredrop through a preferred exemplary HRSG embodiment. The basic engine,controls, packaging, and overall arrangement may be unchanged from theprior art. Therefore, there are no engineering or development costsassociated with this major piece of equipment. This allows the use ofproven technology and helps maintain a high level of power plantreliability. Obviously, GT performance enhancements such as inletchilling, evaporative cooling, and other such methods to increase GToutput may be utilized.

[0431] HRSGs

[0432] The HRSGs from a preferred exemplary embodiment may be smaller,more compact, single pressure level, have controlled heat transfer, andbe optimized for continuous supplemental firing. With one pressure levelversus multi-pressure levels, some preferred exemplary HRSG embodimentsmay be simpler to operate and monitor. Controls may be employed whichcontrol the firing rate, sectional flows, and/or section outlettemperature to provide optimum heat recovery and cycle efficiency for agiven set of operating parameters.

[0433] With the operational flexibility designed into the steam cycle,the GTs will be able to operate at full load over a wide range of powerplant load, providing a more consistent exhaust gas flow to the HRSGsand thus much more efficient performance. Fewer pressure levels, highercycle efficiency, more consistent operation, all lead to betterreliability and lowered O&M costs.

[0434] The need for higher temperature materials or perhaps water-wallconstruction in some preferred exemplary HRSG embodiments will tend toraise the initial cost and also increase the O&M costs. It is doubtfulthat these increased costs will be more than the savings realized fromeliminating other pressure levels, associated controls, and extra heatexchange area.

[0435] HRSGs such as those illustrative of the present inventionteachings currently do not exist in the form as described. However,conventional steam power plant boilers have been built for decades, andthis technology could certainly be applied to some preferred exemplaryHRSG embodiments. In addition, numerous HRSGs have been built withmulti-pressure and single pressure levels, and many have been built withsome degree of supplemental firing (including the higher temperaturewater wall construction). Of all the major components in a preferredexemplary embodiment, this one will require the most engineering anddesign effort. However, as stated previously, the continuously firedHRSG with a single pressure level is a novel concept for thisapplication, but is not beyond technological practice or capability forthose skilled in the art.

[0436] Steam Turbines

[0437] In the prior art, the STs were designed basically by the heatrecovered by the HRSG. On large combined cycle plants, a rule of thumbis that the ST output is approximately 50% of the combined GT output.With supplemental firing this percentage could be increased, but due tothe negative effect on efficiency that was experienced utilizing thehardware, systems, and methods of the prior art, these increases weretypically small. GE informative document GER-3574F (1996), entitled “GECombined-Cycle Product Line and Performance” by David L. Chase, Leroy O.Tomlinson, Thomas L. Davidson, Raub W. Smith, and Chris E. Maslak, inTable 14 indicates that HRSG supplemental firing can increase combinedcycle plant output in the prior art by 28%, but only with an increase inoverall combined cycle heat rate (specific fuel consumption) of 9%.

[0438] The prior art focused on multi-pressure level HRSGs and STs thatused this steam. Consequently, the STs had relatively small HP flows,moderate IP flows due to the addition of IP steam from the HRSG, andrelatively large LP flows due to the further addition of LP steam fromthe HRSG. This yields lower volumetric efficiency for the HP and IPsections of the ST, and potential exhaust end loading problems for theLP section. In addition, the steam cycles themselves were somewhatinefficient, as the IP and LP steam produced by the HRSG had lesspotential to produce work than the HP steam. Finally, the IP and LPsteam flows detrimentally add to both the ST exhaust end loading andalso the heat rejection requirements.

[0439] Due to the volumetric efficiency problems and cost/benefitratios, the inlet pressure ratings for combined cycle plant STs has beenlimited to approximately 1800 psia. As multi-pressure HRSGs have beenemployed, there has been no need for the use of conventional feedwaterheating as there has always been ample heat in the HRSG to provide thisfunction. Thus, the increases in steam cycle efficiency from thisefficiency enhancement feature are not commonly applied.

[0440] The ST utilized in a preferred exemplary embodiment may be largerand more efficient than that from the prior art. The ST utilized in apreferred exemplary embodiment can have a rating of approximately 0.75to 2.25 times (or more) than that of the total GT output. For anequivalent number of GTs and HRSGs capable of firing to 2400° F.,overall combined cycle plant capacity may be increased by a factor of2.00 or more over the prior art. This equates to a ST in several of thepreferred embodiments that can be rated at up to 4.50 times the ratingof the ST from the prior art (a ST that was associated with the sameGTs).

[0441] The ST may be similar to that from the prior art, but will likelyhave an increased inlet pressure rating. In addition, the ST in apreferred exemplary embodiment may utilize extraction steam fedfeedwater heating, which will increase the steam cycle efficiency. Withno IP or LP steam from the HRSG, the steam flow to the HP section of thesteam turbine at rated conditions will be the maximum flow through anysection. This increases HP section volumetric efficiency. From thispoint, steam will be extracted from the ST to various feedwater heaters,fuel preheaters, a smaller steam turbine driven BFP, and/or other plantservices. This operation reduces the exhaust end flow, reducing thepossibility of high exhaust end loading in the ST. All these featuresare typical of a ST that would be used in a conventional steam powerplant.

[0442] Due to its large increase in rating, (from approximately 50% ofGT total capacity to approximately a range of 100% to over 200% of GTtotal capacity), the ST may require larger last stage blades and/or moreLP sections. This represents a relatively low cost addition for capacitycompared to extra GTs, HRSGs, switchgear, transformers, foundations,etc. that would be required in the prior art to provide this extracapacity.

[0443] Other than its larger flow passing capability, higher rating,improved efficiency, and larger blading and/or extra LP section(s), theST may appear similar to a ST in the prior art. It is designed typicallyto extract steam flow from the turbine for conventional feedwaterheating, rather than admit flow to the turbine from the IP and LP HRSGsections. However, the ST would be extremely similar to a ST of similarrating and inlet conditions found in a modern conventional steam powerplant. Therefore, this new combined cycle method allows for the use ofmore conventional, higher efficiency ST hardware and more efficientsteam turbine cycles. This maximizes the bottoming cycle efficiency,vastly increases capacity, and reduces overall combined cycle powerplant size and installed cost, all without a sacrifice in reliability.

[0444] Operation

[0445] With the large amounts of supplemental firing (high ST/GT ratio),and the ability to vary this rate of firing, several preferred exemplaryembodiments become an arrangement where the bottoming cycle is much moreindependent than in the prior art. Due to this phenomenon, and the factthat from an emissions and efficiency standpoint it is best to operatethe GTs at full load, most of the overall combined cycle plant loadvariations in a preferred exemplary embodiment are accomplished byvarying the rate of supplemental firing and subsequently the ST load,while the GTs continue to operate at or near full load. This contrastsfrom the prior art where supplemental firing was utilized to obtain onlyminor increases in plant output during peak operation, and overall plantload control was achieved mainly through load changes on the GTs.

[0446] In several preferred exemplary embodiments, at overall plant fullload, the GTs will be at full load, and either the HRSG will havereached its firing temperature limit, or the ST will have reached itsinlet pressure limit. From this point, as plant load is reduced,supplemental firing is reduced, steam production is reduced, andsubsequently the ST load is also reduced. This process of load reductioncontinues until adequate flows can no longer be maintained in the HRSG.

[0447] Once adequate flows can no longer be maintained in the HRSG, theST and/or HRSG will reach an operational limit. At this point it will benecessary to decrease load on a GT or GTs. As the total GT load isreduced, ST load can be increased to meet system load. Refer to FIG. 43for a suggested mode of operation with multiple GTs. This control methodmay be used to reduce load from overall plant full load down to the HRSGand/or ST low limit by varying the rate of supplemental firing only, andallowing the GTs to operate at full load. Once at this low limit, one GTcan be unloaded, and its HRSG will begin to produce less steam.Concurrently, the remaining GT can remain at or near fall load, and itsHRSG can increase its rate of supplemental firing. This results in moresteam to the ST. The net result is a transitional zone of operationwhere one GT is reduced in load while the ST compensates for most ofthis load reduction. After reducing overall plant load sufficiently topass through this transitional zone of operation, one GT will be takenout of service (shut down), and the remaining HRSGs will be supplementalfiring at high rates and the ST will be operating at a much higher loadthan at the upper end of the transitional zone. This scheme of operationallows the GTs to remain at or near full load through a large range ofthe overall plant's expected output (approximately 50 to 100% of plantrating) with only a narrow band of operation in the transitional zonewhere one GT is brought from full load to an out-of-service condition.For FIG. 43, this transitional zone of operation is between 70% and 80%of plant load.

[0448] An exemplary embodiment of a control structure implementing theabove procedures is illustrated conceptually in the flowcharts of FIGS.16-19. Discussion of this embodiment is detailed later in this document.

[0449] Performance

[0450] Since the rate of supplemental firing is large compared to theprior art, the ST capability is greatly increased. By utilizing an HRSGcapable of 2400° F. inlet temperatures, the ST can be designed (forexample) at its rated point to be approximately 2.25 times the output ofall the GTs combined. This is substantially more than a ST from theprior art, as in these applications, ST rated output was typically inthe range of 0.4 to 0.6 times the output of all the GTs. This greatlyincreases the capacity of the power plant, as the ST is now capable ofratings that are up to 4.50 times that of the ST in the prior art. Also,as previously mentioned, the operational flexibility afforded by thisarrangement allows for operation of the GTs at full load over a widerange of overall plant output. This increases the plant's part loadefficiency and lowers NOX emission levels for GTs which typicallydemonstrate increased NOX emissions at part load operation.

[0451] With this large increase in capacity over the prior art, theadded flexibility, and lowered cost per kW of capacity, this example ofa preferred exemplary embodiment combined cycle plant is more adept bothoperationally and economically to provide the temporary powerrequirements of seasonal peak loads. In addition, small operationalvariables (like the isolation of feedwater heaters or operation with theHP inlet pressure at 5% over rated) will allow this example of apreferred exemplary embodiment to attain even greater capacity thanrated, but at a slight cost in efficiency. Since seasonal peaks may havedurations that last for only a matter of days each year, this is aninexpensive method to generate more power during peak periods (which maybe sold at very high rates) for minimal cost. The increased revenue isenvisioned to more than compensate for the inefficiencies and theincreased fuel costs incurred during these temporary peak loadingconditions, thus making this an economically advantageous alternativefor plant designers and electric utilities. As reported in POWERMAGAZINE, (ISSN 0032-5929, March/April 1999, page 14):

[0452] “Reserve margins are down nationwide . . . Last summer's Midwestprice spikes, up to [US]$7000/MWh [(US$7.00/kWh)], garnered most of thepress coverage, but spikes of [US]$6000/MWh [(US$6.00/kWh)], alsooccurred in Alberta . . .

[0453] Although it has been stated in the prior art that supplementalfiring decreases overall combined cycle thermal efficiency, this exampleof a preferred exemplary embodiment has shown this assumption to beincorrect. By utilizing the fuel added in supplemental firing to notonly add heat, but upgrade the bottoming cycle efficiency, it ispossible to meet or exceed prior art overall combined cycleefficiencies. This is accomplished through the use of higher inlet steampressures, larger more efficient STs, the conversion of lower pressuresteam utilized in the prior art to high-pressure steam, and the use ofconventional feedwater heating. Part load operation is also improved asthe GTs in this example of a preferred exemplary embodiment will operateat full load (where they are most efficient) for a vast majority oftheir operation (neglecting the time when they may be out of service).

[0454] Part Load Performance

[0455] As system load is reduced, the combined cycle plant load must bereduced to meet the electrical system demand. In the prior art, this wasaccomplished by a reduction in load on the GTs. This mode of loadcontrol causes a decay in the GT efficiency as well as the overallcombined cycle plant efficiency. With several of the preferredembodiments of the present invention, however, load control isaccomplished more through the variation of the amount of supplementalfiring. In this manner, the GTs remain at or near full power where theyare the most efficient and have the lowest emissions. The bulk of theload modulation is then accomplished by a reduction in the amount ofsteam production and a subsequent reduction in the output of the ST.This mode of operation provides for improved part load efficiency forthe overall combined cycle plant, as well as a reduction in maintenanceon the GTs as a result of the reduction in thermal cycling operation (GTinternal temperatures typically vary with changes in GT loading).

[0456]FIG. 33 indicates some part load efficiencies that can be expectedfrom conventional combined cycle power plants in the prior art and alsothose that can be attained with several of the preferred embodiments ofthe present invention. As can be seen from these curves, the prior artcombined cycles continually degrade from their optimum performance asload is reduced from 100%. However, several of the preferred embodimentsof the present invention actually experience an increase in efficiencyas load is initially reduced from 100% before it begins to degrade belowabout 80%. This part load efficiency profile for several of thepreferred embodiments of the present invention provide for substantialfuel savings as compared to the conventional combined cycle in the priorart.

[0457] Peak Load Performance

[0458] The present invention is particularly well suited for providingpower at periods of peak load. During these periods, the output of apreferred embodiment combined cycle power plant may be temporarilyextended beyond its nominal rated load. As mentioned previously, thistemporary extension beyond rated power plant load may provide anenormous economic benefit, as peak power can sell for hundreds of timesthe normal price of non-peak generated electric power. Therefore, thereis a strong incentive for power plant owners to generate this power. Aspreviously mentioned, the prior art has addressed this problem byutilizing supplemental firing in the HRSG. Not only does this reduce theconventional combined cycle efficiency in the prior art at peak loads,but also due to the need for added ST capacity the base prior artcombined cycle efficiency is also reduced at non-peak loads as well (STis already at part load with no supplemental firing). Thus, the abilityto extend the peak power rating of conventional combined cycle powerplants comes with a detriment to the overall plant efficiency at allplant load operating points.

[0459] Since ST capacity can be increased through greater mass flow,techniques that increase steam flow through the ST will normallyincrease overall ST output. Since the present invention teaches apredominantly Rankine cycle combined cycle, and as such, increases inthe ST output affect a wider variation in the overall combined cyclepower plant capacity. Therefore, this effect to ST output is much moreeffective than in the primarily GT-based combined cycle power plants astaught by the prior art.

[0460] Note in the following table that as the pressure is increased acorresponding increase in steam flow takes place. If this pressureincrease is coupled with a corresponding decrease in inlet steamtemperature, further increases in mass flow are attainable. Inconjunction with this, isolation of feedwater heaters will serve todirect more steam flow to the exhaust of the ST, further increasing SToutput. Unlike the prior art, this method provides the ability to extendthe peak power rating of combined cycle power plants implementing thepresent invention without incurring a detriment to the overall plantefficiency at non-peak plant load operating points. Peak Power ExtensionExample Inlet Tem- Mass/ Inlet perature Specific Steam Volume Steam FlowPressure (degrees Volume Flow Flow Increase (psia) F) (ft³/lb) (lb/hr)(ft³/hr) (new/old) 2400 1050 0.338245 2000000 676490 Baseline 2520 10500.320349 2111730 676490 1.055865 2520 1000 0.304021 2225143 6764901.112571 2520 950 0.286872 2358162 676490 1.179081 2640 1000 0.2882362346998 676490 1.173499 2640 950 0.271554 2491183 676490 1.245591

[0461] Cost

[0462] A substantial advantage to this exemplary preferred embodiment isthe cost savings. As mentioned previously, a plant with HRSGs designedfor up to 2400° F. inlet temperature through supplemental firing caneasily have a ST rated 2.0 times the total GT capacity. Therefore, totalplant output is 3.0 (2.0ST+1.0 GT) times the GT capacity. A combinedcycle plant from the prior art would have a ST rated at approximately0.5 times the total GT capacity. Therefore, the capacity ratio isessentially (3.0/1.5)=2.0. In other words, the combined cycle plant fromthis preferred exemplary embodiment will have 100% greater capacity thanthe prior art. An example of this trend is demonstrated in FIG. 39,which is a heat balance for a 1040 MW exemplary preferred embodimentutilizing two (2) industry standard GE model PG7241FA GTs and a largeST. FIG. 22 illustrates a combined cycle from the prior art utilizingthe same quantity and model of GTs and the standard smaller ST,nominally rated at 520 MW.

[0463] This means that to provide capacity equal to that from thisexample, a combined cycle plant from the prior art would need to add100% more equipment. This means more GTs, another ST, more HRSGs,switchgears, transformers, and all the necessary systems and real estaterequired to support this equipment. This will serve to raise the plantinstalled cost by essentially 100%.

[0464] In terms of 1999 dollars, a modem high efficiency large combinedcycle power plant could be installed for approximately US$450 per kW ofcapacity. Therefore, a 720 MW plant (720,000 kW) would cost US$324million to construct. If this plant were to be expanded to 1050 MW, theinstalled cost would climb to US$472 million. In contrast, the presentinvention teaches that it is possible to use less equipment to affectthis expansion, thus decreasing the cost per kW of rated plant capacity.

[0465] Retrofits

[0466] Another prime application for this example of a preferredexemplary embodiment is in retrofit applications of existing plants.Many steam-powered plants in existence today will produce expensivepower compared to the highly efficient combined cycle plants discussedherein. With electrical deregulation on the horizon, it will beimperative that power producers be competitive. Therefore, technologythat will help existing steam plants compete with new combined cycleplants is needed.

[0467] Since this example of a preferred exemplary embodiment operates(predominantly) on a single pressure level, utilizes higher steampressures that are typical for STs found in conventional steam plants,has a higher ST/GT output ratio, and provides for a compact design, itis ideally suited for retrofit applications of existing steam powerplants. With a preferred exemplary embodiment, large steam plants couldactually bypass their existing boilers and utilize steam directly fromthe HRSGs. This increases cycle efficiency and (in many cases) wouldreduce plant emissions drastically. This could be accomplished using theexisting ST, condenser, and other infrastructure already in the existingplant. This would provide the owners with a highly efficient combinedcycle plant with reduced capital investment and minimal real estaterequirements.

[0468] Exemplary Preferred Embodiments—Typical Configuration

[0469] Overview

[0470] The configuration of several of the preferred embodiments issimilar to the prior art, in that GTs and HRSGs are utilized to producepower and convert exhaust gas heat into steam. However, several of thepreferred embodiments will utilize a continuously fired HRSG thatproduces significantly more steam, and do so at a single pressure level(or primarily a single pressure level). This higher quantity (andtypically higher pressure) steam drives a ST that is much larger incomparison to the ST in the prior art that was associated with the sameGTs.

[0471] Due to the large feedwater flows, feedwater will be heated in theHRSGs as well as in a separate feedwater heating loop which utilizesconventional ST extraction steam fed feedwater heaters. Fuel gas heaterswill also be employed to improve cycle efficiency.

[0472] Embodiment of FIG. 9

[0473] Design

[0474] Refer to FIG. 9 for a schematic representation of one exemplarypreferred power plant embodiment utilizing the teachings of the presentinvention. The GTs (920) each exhaust into their respective HRSGs anddrive their respective generator (921). These exhaust gases producesteam in the HRSG that subsequently produces power in the ST and whichis ultimately condensed in the condenser (939).

[0475] Feedwater Heating—HRSG Feedwater Heating Loop

[0476] Condensate from the condenser (939) goes to the LP-BFP (930)where it is pumped to an intermediate discharge pressure. From here, theLP feedwater control valve (960) maintains an optimum flow through theLP economizer (901) while diverting the excess feedwater flow to theconventional feedwater heater(s) (933). Flow exiting the LP economizercontinues to the HP-BFP (931) and is pressurized to a pressure that isequal to inlet steam pressure plus an allowance for pressure drops inthe system. From here it flows through the HP economizer (902) and(903). Some feedwater flow, however, after exiting the LP economizer, isdiverted through the feedwater balancing valve (967) so as to maintainan optimum flow through the HP economizer sections (902) and (903). Thediverted feedwater that passes through the feedwater balancing valve(967) combines with the feedwater exiting feedwater heater (933). Thiscombined flow now continues to the second HP-BFP (932) where it ispressurized to a pressure similar to that of HP-BFP (931). The divertedfeedwater flow exiting HP-BFP (932) goes to feedwater heater(s) (934).The feedwater exits feedwater heater(s) (934) and combines with thefeedwater flow exiting the HP economizer. This flow is now available atdesuperheating lines (950) and (951), while the bulk of the flowcontinues to the evaporator section (904).

[0477] Evaporator

[0478] In the evaporator section, the feedwater is boiled into steam andtravels to the superheater section (905). If the superheated steam istoo hot, condensate is sprayed through line (950) into the superheatersupply line to control the HP turbine section (935) inlet temperature tothe desired temperature. Steam expands in the HP turbine section down tothe exhaust point, and becomes known as cold reheat steam. The coldreheat steam continues to the reheater section (906) in the HRSG.

[0479] Reheater

[0480] On its way to the reheater section, some steam passes throughnon-return valve (964) to line (954). This steam travels to thefeedwater heater (934), which preheats the feedwater flowing throughsame. The condensed steam from this feedwater heater cascades down tothe inlet of the HP-BFP (932).

[0481] The cold reheat steam from the HP turbine section exhaust nowtravels through the reheater section of the HRSG (906) for return to theIP section of the ST. If its temperature is too high, condensate issprayed through line (951) into the reheater supply line to control theIP turbine section (936) inlet temperature to the desired temperature.Steam expands in the IP turbine section down to its exhaust point, andbecomes known at this point as crossover steam.

[0482] Crossover Steam

[0483] The crossover steam continues to the LP sections (937) of the ST.On its way to the LP section, some crossover steam is diverted throughnon-return valve (965) to line (955). This steam travels to feedwaterheater (933), which preheats the feedwater flowing through same. Thecondensed steam from this feedwater heater flows to the outlet of thecondenser.

[0484] Steam expands in the LP turbine sections and exhaust into thecondenser (939). Shaft horsepower produced in the ST drives thegenerator (938), which produces electrical power.

[0485] Note that in this example cold reheat and crossover steam is usedto provide extraction steam to the feedwater heaters. Although these aretraditional points for the supply of this steam, this does not precludethe utilization of extraction steam from any practival point on the STto provide this function.

[0486] Low Load Operation

[0487] For operation at low loads, there is insufficient HP steam flowto maintain optimum levels of feedwater through the HRSG. In this modeof operation, valves (960) and (967) are closed. With no feedwater flowto remove heat, all extraction lines (954, 955) pass no flow. Allfeedwater flow, therefore, passes through the HRSG as the parallelfeedwater loop is closed off.

[0488] As load is decreased from this point by a reduction in steam flow(reduction in supplemental firing), the feedwater flow through the HRSGis no longer sufficient to absorb the exhaust gas heat and yet stillmaintain optimum exhaust gas temperature. Therefore, operation belowthis point will result in increased exhaust gas temperatures and lowercombined cycle efficiency. At this point, the design engineer will needto evaluate performance parameters and determine if it is moreeconomical at this point of operation to reduce load of the GTs, orcontinue modulating supplemental firing rates and allowing the HRSGexhaust gas temperature to increase. At some point of reduced load,however, it will become economically favorable to reduced load on theGTs.

[0489] Embodiment of FIG. 15

[0490] Design

[0491] Refer to FIG. 15 for a schematic representation of anotherexemplary preferred power plant embodiment utilizing the teachings ofthe present invention. The GTs (1520) each exhaust into their respectiveHRSGs (1509) and drive their respective generator (1521). These exhaustgases produce steam in the HRSG that subsequently produces power in theST and which is ultimately condensed in the condenser (1595).

[0492] Feedwater Heating—HRSG Feedwater Heating Loop

[0493] Condensate from the condenser (1595) goes to the LP-BFP (1530)where it is pumped to its discharge pressure. From there, the LPfeedwater control valve (1560) maintains an optimum flow through the LPeconomizer (1501) while diverting the excess feedwater flow to the firstof a series of conventional feedwater heaters (1533). Flow exiting theLP economizer continues to the HP-BFP (1531) and is pressurized. Fromhere it flow through the HP economizer (1502). However, after exitingthe LP economizer some feedwater flow is diverted through the feedwaterbalancing valve (1561) so as to maintain an optimum flow through the HPeconomizer section (1502). In addition, some flow is diverted to thefuel gas heater (1575) through line (1571). After pre-heating the fuelgas, this flow is returned to the inlet of the LP-BFP (1530) via line(1572). The remaining feedwater continues to the HP economizer, and flowexiting the HP economizer combines with the feedwater flow exiting thefinal feedwater heater (1537). This flow is now available atdesuperheating valves (1510) and (1511), while the bulk of the flowcontinues to the evaporator section (1504).

[0494] Feedwater Heating—Conventional Feedwater Heating Loop

[0495] In the parallel feedwater heating loop, feedwater proceedsthrough the first feedwater heater (1533) where it is heated. This flowthen travels through the second and third feedwater heaters (1534) and(1535) respectively. At the exit of feedwater heater (1535), flowdiverted from the HRSG parallel loop through the feedwater balancingvalve (1561) combines with this feedwater and continues to a HP-BFP(1532) where it is pressurized. From here it travels through the fourthand fifth feedwater heaters (1536) and (1537) respectively. Thefeedwater from this heating loop now combines with the feedwater fromthe HRSG parallel loop and is fed to the evaporator section (1504) ofthe HRSG (minus flow required by the desuperheating valves (1510) and(1511)).

[0496] Evaporator

[0497] In the evaporator section, the feedwater is boiled into steam andtravels to the superheater section (1505). If the superheated steam istoo hot, desuperheating valve (1510) modulates to spray condensate fromthe desuperheating line (1550) into the superheater supply line andcontrol the HP turbine section (1590) inlet temperature. Steam expandsin the HP turbine section until reaching the first extraction where asmall portion of the steam is removed from the turbine throughnon-return valve (1568) to line (1558). This steam is fed to the fifthfeedwater heater (1537) which preheats the feedwater flowing throughsame. The condensed steam from the fifth feedwater heater cascades downto the fourth feedwater heater (1536). The steam in the HP section ofthe ST (1590) that is not extracted continues to the section exit point,and becomes known as cold reheat steam. The cold reheat steam continuesto the reheater section (1506) in the HRSG.

[0498] Reheater

[0499] On its way to the reheater section, some steam (secondextraction) passes through non-return valve (1564) to line (1554). Thissteam travels to the fourth feedwater heater (1536) which preheats thefeedwater flowing through same. The condensed steam from the fourthfeedwater heater cascades down to the inlet of the HP-BFP (1532).

[0500] The cold reheat steam now travels through the reheater section ofthe HRSG for return to the IP section of the ST. If its temperature istoo high, desuperheating valve (1511) modulates to spray condensate fromthe desuperheating line (1551) into the reheater supply line and controlthe IP turbine section (1591) inlet temperature. Steam expands in the IPturbine section until reaching the third extraction where a smallportion of the steam is removed from the turbine through non-returnvalve (1567) to line (1557). This steam is fed to the third feedwaterheater (1535) which preheats the feedwater flowing through same. Thecondensed steam from the third feedwater cascades down to the 2^(nd)feedwater heater. The steam in the IP section of the ST (1591) that isnot extracted continues to the section exit point, and becomes known ascrossover steam.

[0501] Crossover Steam

[0502] The crossover steam continues to the LP sections (1592) and(1593) of the ST. On its way to the LP section, some steam (fourthextraction) is diverted through non-return valve (1565) to line (1555).This steam travels to the second feedwater heater (1534) which preheatsthe feedwater flowing through same. The condensed steam from the secondfeedwater heater cascades down to the first feedwater heater (1533).

[0503] Steam expands in the LP turbine sections until reaching the fifthextraction where a small portion of the steam is removed from theturbine through non-return valve (1569) to line (1559). This steam isfed to the first feedwater heater (1533) which preheats the feedwaterflowing through same. The condensed steam from the first feedwaterheater is returned via line (1512) to the inlet of the LP-BFP (1530).

[0504] The steam in the LP sections of the ST (1592, 1593) that is notextracted continues through the section to exit at the condenser (1595).Shaft horsepower produced in the ST drives the generator (1594) whichproduces electrical power.

[0505] Low Load Operation

[0506] For operation at low loads, there is insufficient HP steam flow(thus low flows of condensate from condenser) to maintain optimum levelsof feedwater through the HRSG. In this mode of operation, valves (1560)and (1561) are closed. With no feedwater flow to remove heat, allextraction lines (1558, 1554, 1557, 1555, 1559) pass no flow. Allfeedwater flow, therefore, passes through the HRSG as the parallelfeedwater loop is closed off.

[0507] As load is decreased from this point by a reduction in steam flow(reduction in supplemental firing), the feedwater flow through the HRSGis no longer sufficient to absorb the exhaust gas heat and yet stillmaintain optimum exhaust gas temperature. Therefore, operation belowthis point will result in increased exhaust gas temperatures and lowercombined cycle efficiency. At this point, the design engineer will needto evaluate performance parameters and determine if it is moreeconomical at this point of operation to reduce load on the GTs, orcontinue modulating supplemental firing rates and allowing the HRSGexhaust gas temperature to increase. At some point of reduced load,however, it will become economically favorable to reduce load on theGTs.

[0508] Exemplary Preferred Embodiment—725 MW Power Plant

[0509] Overview

[0510] As an example of another preferred exemplary embodiment, a 725 MWnominal capacity combined cycle power plant design will be examined.This exemplary power plant will utilize two (2) GE Model PG7241(FA) GTs.These GTs will each exhaust into its own single pressure HRSG designedfor 2400 psia operation. A nominal 400 MW reheat ST will be usedexhausting to a once through condenser operating at 1.2 inches HgA(inches of mercury absolute) exhaust pressure. Due to the largefeedwater flows, feedwater will be heated in the HRSGs as well as in aseparate feedwater heating loop which utilizes conventional STextraction steam fed feedwater heaters. Fuel gas heaters will also beemployed to improve cycle efficiency.

[0511] Design

[0512] The GE GT design is rated 170,770 kW based upon ISO conditions,with a 3.0 inches of H₂O inlet air pressure drop and 10.0 inches H₂Oexhaust gas pressure drop through the HRSG. Total GT output is therefore341,540 kW. Refer to FIG. 35 for a schematic representation of thisexemplary power plant. The numbers indicated at various points along theprocess correspond to “point” numbers tabulated in FIGS. 36, 37, and 38.The data corresponding to the “point” numbers tabulated in FIG. 36, FIG.37, and FIG. 38 identifies the pressure, temperature, enthalpy, and flowat the corresponding “point” . This overall information contained inFIGS. 35-38 represents what is termed a “heat balance”, which is anoverall energy and mass balance for the cycle. Note for this example,deaeration is completed in the condenser.

[0513] Layout

[0514]FIG. 26 illustrates the physical plant layout of this example ofseveral of the preferred embodiments. Note that it is extremely similarto the GE S207FA combined cycle power plant in the prior art, shown inFIG. 22. The most noticeable difference between the two layouts is theconfiguration of the ST. In the prior art, the ST has a relativelyunderutilized HP/IP section, and one LP section. In several of thepreferred embodiments, the HP/IP section is similar to the prior art,but has considerably increased volumetric flow. To efficiently use thehigher steam flows at lower pressure, a second LP section is shown.However, this second section may not be required, depending upon theeconomic evaluation.

[0515] Comparison to Prior Art

[0516]FIG. 22 and FIG. 24 are layouts of the GE S207FA combined cycleand the Westinghouse 2X1 501G combined cycle power plants respectively.The GE facility requires approximately 2.3 acres of real estate whilethe Westinghouse facility requires approximately 3.3 acres. The powerdensity is nearly the same for these two options at 220 MW per acre.Several of the preferred embodiments, however, can be designed as shownin FIG. 26 to be 725 MW as in the example, which is 315 MW per acre, orit can be designed for up to 1050 MW (see FIG. 29) which is 455 MW peracre. This allows for the production of significantly more power withonly a given amount of real estate. This factor is advantageous for newconstruction, but will also be especially appreciated for retrofit ofexisting plants where real estate comes at a premium.

[0517] Besides the premium for real estate, the combined cycles in theprior art are also more expensive from a fuel consumption, capital cost,and maintenance perspective. FIG. 23 and FIG. 25 are economic pro formafor the GE S207FA combined cycle and the Westinghouse 2X1 501G combinedcycle power plants respectively. These figures tabulate the annual costsfor fuel, capital, and maintenance for each power plant. FIG. 27 is theeconomic pro forma for an exemplary preferred embodiment of the presentinvention. Note that each individual cost for fuel, capital, andmaintenance is less than the each individual cost for combined cyclepower plants from the prior art. Therefore, the cost to produceelectricity is reduced in all major cost categories by several of thepreferred embodiments.

[0518] Exemplary Preferred Embodiment—Supercritical Steam Conditions

[0519] Overview

[0520] As another example of a preferred exemplary embodiment, a 1040 MWnominal capacity combined cycle power plant design utilizingultrasupercritical steam conditions with elevated steam temperatureswill be examined. This exemplary power plant will utilize two (2) GEModel PG7241(FA) GTs. These GTs will each exhaust into its own singlepressure HRSGs designed for 3860 psia operation. A nominal 730 MW doublereheat ST will be used exhausting to a once through condenser operatingat 1.2 inches H_(g)A (inches of mercury absolute) exhaust pressure. Dueto the large feedwater flows, feedwater will be heated in the HRSGs aswell as in a separate feedwater heating loop which utilizes conventionalST extraction steam fed feedwater heaters. Fuel gas heaters will also beemployed to improve cycle efficiency.

[0521] Design

[0522] The GE GT design is rated 168,815 kW based upon ISO conditions,with a 3 inches of H20 inlet air pressure drop and a 10.0 inches H2Oexhaust gas pressure drop through the HRSG. Total GT output is therefore341,540 kW. Refer to FIG. 39 for a schematic representation of thisexemplary power plant. The numbers indicated at various points along theprocess correspond to “point” numbers tabulated in FIGS. 40, 41, and 42.The remaining data corresponding to the “point” numbers tabulated inFIG. 40, FIG. 41, and FIG. 42 identifies the pressure, temperature,enthalpy, and flow at the corresponding “point”. This “heat balance” isan overall energy and mass balance for the cycle. Note for this example,deaeration is completed in the condenser.

[0523] Comparison to Prior Art

[0524] The elevated steam temperatures (1112° F.) and pressures (3860psig) are indicative of those used in advanced steam cycles, sometimesreferred to as ultrasupercritical. Refer to the informative documententitled “Steam Turbines for Ultrasupercritical Power Plants” by KlausM. Retzlaff and W. Anthony Ruegger (General Electric Reference GER-3945,1996) for information on ultrasupercritical steam turbines and theircycles. Note that at an exhaust temperature of 1123° F., the industrystandard General Electric (GE) Model PG7241(FA) Gas Turbine does nothave sufficient high temperature exhaust energy to produce these steamtemperatures at the required flows. Therefore, such conditions were noteven available in the prior art.

[0525] The supercritical steam power plant of the preferred embodimentof the present invention is similar to the subcritical steam power plantof the preferred embodiment, with the primary difference being improvedefficiency. Greater steam pressures, higher steam temperatures, and theuse of the second reheat provides the added efficiency for thisapplication. Note that with these steam conditions, and even with thelarge extension in capacity (100%) the combined cycle efficiency for thepreferred embodiment of the present invention approaches that of theprior art with the same technology GTs (6229 BTU/kWh versus 6040BTU/kWh).

[0526] However, efficiency is only one part of the economic equation.The other major costs, capital expenditure and maintenance, will begreater with the supercritical preferred embodiment versus subcritical.Therefore, as previously discussed, a total economic analysis must becompleted to determine the optimum arrangement for an individualpreferred embodiment combined cycle power plant. In general, when fuelcosts are high, supercritical applications will become the economicoptimum, and when fuel costs are low, subcritical applications will bepreferred.

Power Plant Load Profile

[0527] Dispatched Power Plants

[0528] As previously discussed, to maintain a constant frequency ofpower (60 Hz in the US), the power produced by all power plantsconnected to the grid must equal the power being consumed by the userson the grid. Therefore, power plants have their output “dispatched”, orcontrolled by the Power Pool to meet the system demand.

[0529] As a result of being dispatched, most power plants will spendvery little of their operational time at rated output. Instead ofoperating at full rated capacity, most power plants will operate at someintermediate load and share the system load with all other power plantsconnected to the grid. This statistic may be visually confirmed byinspecting the load duration curve of FIG. 31B, which represents atypical long-term distribution of utilized plant load versus percentageof time. Note that using this long-term data, most power plants willoperate at peak load less than 10% of the time, and will be atintermediate load levels for 70% of the time.

[0530]FIG. 31A provides typical hourly load data for the South AtlanticRegion of the U.S. over a 24-hour period. As can be seen from this data,the peak load of 62,000 megawatts (MW) for the day is substantiallyhigher than the low of 40,000 megawatts. In addition, the total systemcapacity is likely higher than 62,000 MW, perhaps 70,000 MW (70gigawatts, GW). This means that except for seasonal peaks (i.e. hotsummer days), even during non-seasonal peak hours, many power plants arenot operated at rated capacity. Therefore, dispatched power plants canexpect to see large load variations and potentially spend only a matterof hours annually at rated capacity.

[0531] To determine a typical conservative load profile, the data fromFIG. 31A was blocked into segments. The periods when the load was above60 GW was determined to be peak operation. The periods of operationbetween 50 and 60 GW was considered to be intermediate power operation,and periods below 50 GW were considered to be night operation. Thisprofile was considered to be an average weekday. For weekends, 8 hoursper day was considered intermediate, while the remainder was taken to benight operation (using weekday averages for intermediate and night poweron weekends). FIG. 32 provides the details of these calculations.Utilizing the data calculated in FIG. 32, a typical load profile to beused for comparison purposes is as follows: Period Average PlantCapacity (%) Hours Per Week Night 60 77 Intermediate 80 71 Peak 100 20

[0532] Note that although the capacity per FIG. 32 for peak is only87.86%, this number has been adjusted to 100.00% for discussionpurposes. The night and intermediate capacity numbers have been adjustedby less than 1% from the values in FIG. 32, and are adjusted downward tocompensate for the upward adjustment to peak operation.

[0533] Exemplary Power Plant Load Profile

[0534] Utilizing the data from the above table, the calculated loadprofile can be used for the purpose of determining an annual capacityfactor and quantity of fuel consumed for a given combined cycle powerplant, based upon part load operation data in FIG. 33. It is significantto note from the table above and FIG. 31B that the plant efficiencyusing the prior art technology will rarely (if ever) reach optimumeconomic performance. In contrast, the present invention embodiments asillustrated in FIG. 33 will always be more optimal than the prior artconfigurations.

ECONOMICS OF THE PRESENT INVENTION

[0535] Economic Considerations

[0536] The costs for operating a combined cycle power plant are varied.However, the three largest costs for the power plant operators typicallyare fuel, capital cost (debt), and maintenance. These three costsconstitute the major portion of the cost (expressed in $/kWh) to produceelectricity at large combined cycle power plants. Some of the minorcosts include payroll for the operations staff, taxes, insurance,license fees, and other miscellaneous expenses. For an economiccomparison of several of the preferred embodiments of the presentinvention to the prior art, focus will be on the three major expenses:fuel, debt, and maintenance.

[0537] Fuel Costs

[0538] The largest cost that typically is incurred by a large, modem,combined cycle facility is the cost of fuel. Whether the fuel is naturalgas, fuel oil, or some other combustible fuel, the combined cyclefacility must consume large quantities of fuel to produce largequantities of electricity. In essence, a power plant actually convertsenergy in one form (raw fuel), into energy of another form(electricity). Therefore, since the function of a power plant is toperform this conversion process, the efficiency of this conversionprocess is the key to the power plant's economic success.

[0539] Prior art combined cycle power plants have efficiencies in thegeneral range of 48% (LHV) for an older design such as a GE S106Bcombined cycle up to 60% (LHV) for the proposed GE S107H advanced cyclewhich has not yet seen commercial service. These efficiencies are basedupon the lower heating value (LHV) of the fuel. However, theseefficiencies are for full load operation, and as noted in FIG. 31A andFIG. 31B, most power plants actually spend little time at full load. Forpart load operation, FIG. 8 provides an indication of the efficiencyloss that can be expected at reduced loads for combined cycle powerplants in the prior art. Utilizing this data, FIG. 33 illustrates thedramatic improvement in part load efficiency that is realized by severalof the preferred embodiments of the present invention as compared to thecombined cycle in the prior art (here a lower heat rate indicates moreoptimal performance). This part load efficiency improvement, along withimproved efficiency at full load, enables several of the preferredembodiments of the present invention to be more economical than theprior art in terms of fuel consumption.

[0540] Based upon the load profile in FIG. 32, and utilizing the heatrate (efficiency) data from FIG. 33, FIG. 34 tabulates the annual fuelcosts for this exemplary combined cycle power plant of the preferredembodiment versus current state-of-the-art combined cycle power plantsin the prior art. In either case, many of the exemplary combined cyclepower plants of the preferred embodiment use less fuel on an annualbasis than either of the prior art combined cycles.

[0541] Capital Costs

[0542] Next to fuel costs, the most significant cost for a new combinedcycle power facility is capital cost. This is the amount of moneyrequired to service the debt (loan payments). Although plant efficiencyis important, the overall cost of the power plant is also an importanteconomic consideration. As discussed prior, just as the economics ofsmall portions of the combined cycle plant must be evaluated (i.e.larger ST exhaust sections), the economics of the overall combined cyclepower plant must also be evaluated. Minor decreases in plant heat rate(minor increase in efficiency) must not be more than offset by increasesin capital cost. Therefore, the power plant developers and engineersstrive to construct the best economic alternative that is available.

[0543] Due to its higher power density, utilization of less equipment,and reduced construction costs, several of the preferred embodiments ofthe present invention have significantly lower capital costs (up to a30% reduction) than combined cycles in the prior art. Again, FIG. 34tabulates the capital costs for this exemplary combined cycle powerplant of the preferred embodiment versus current state-of-the-artcombined cycle power plants in the prior art. In either case, manyexemplary combined cycle power plants of the preferred embodimentrequire significantly less capital than either of the prior art combinedcycles.

[0544] Maintenance Costs

[0545] Another large expense for power plant owners is average annualmaintenance costs, especially maintenance costs for the large pieces ofequipment. For a large 725 MW plant in the prior art, as shown in theexample, these costs can exceed $10 million annually. Therefore, powerplants with reduced maintenance costs are economically advantageous.

[0546] By utilizing a high power density design which reduces the amountof major equipment, and by utilizing low maintenance STs as the majorpower producing machines instead of high maintenance GTs, several of thepreferred embodiments of E the present invention have appreciably lowermaintenance costs than combined cycles in the prior art. In FIG. 34maintenance costs for this exemplary combined cycle power plant of thepreferred embodiment versus current state-of-the-art combined cyclepower plants in the prior art are tabulated. In either case, theexemplary combined cycle power plant of the preferred embodiment is lessmaintenance intensive than either of the prior art combined cycles.

[0547] Overall Cost Comparison

[0548]FIG. 34 provides an economic comparison of the exemplary combinedcycle power plant of a preferred embodiment of the present invention incontrast to state-of-the-art combined cycle power plants in the priorart. As can be seen from the data, this exemplary combined cycle powerplant of the preferred embodiment is less expensive to operate thancombined cycles in the prior art in all three of the major costcategories: fuel, capital expenditures, and maintenance.

[0549] In addition, compared to the Westinghouse 2X1 501G combined cyclepower plant, NOX emissions are reduced by a factor of more than three,or by approximately 180 tons/yr. For a 20-year plant operational life,the exemplary combined cycle power plant of the illustrated preferredembodiment saves US$469 million as compared to the Westinghouse model501 G combined cycle from the prior art. These savings are more than theinitial plant construction costs of US$340 million for the Westinghouse2X1 501G combined cycle power plant, and represent a significanteconomic advantage for power producers in a deregulated, competitiveenvironment.

OPERATION OF THE PRESENT INVENTION

[0550] Exemplary HRSG Control Method

[0551] Due to the unique arrangement of equipment, the use of apredominantly single pressure level HRSG, and the need to optimize heatrecovery, an exemplary control system to meet these objectives isillustrated in FIG. 16. The control system is exemplary of a combinedcycle described in the preferred embodiments illustrated in FIG. 9 andFIG. 15, although it may have a wide application to other embodiments ofthe present invention. There is one HRSG for each GT in this example.Note that this is an example of an HRSG control system for thisparticular application, and is a demonstration of the principles in flowmanagement, optimum heat transfer, and integration of HRSG and feedwaterheating loops. For other applications, this arrangement could bemodified for the particular circumstances. However, many of theprinciples outlined in this control schematic would be employed.

[0552] In FIG. 16, the control begins at (1601) and continues to processblock (1602) where the loop control begins. Control then flows toprocess block (1603). At this point the controller examines inputs fromprocess block (1611) which include ambient temperature and GT load (inparticular, the GT exhausting into the HRSG in this control loop). Basedupon a characteristic curve programmed into the software, the controllerdetermines the GT exhaust flow.

[0553] Utilizing the DCS inputs for ST required steam flow and steamflow already being produced by the other HRSGs, at process block (1604)the controller calculates its required steam flow as the ST requiredflow minus flow from other HRSGs. Control proceeds to decision block(1605) and compares the HRSG required steam flow to the optimum flow forthe HP economizer.

[0554] If the power plant is operating at reduced load, control flows toprocess block (1606). At this point of operation, there is less than theoptimum HP economizer flow required from the HRSG. Therefore, more heatwill be available in the GT exhaust gases than can be recovered in theHRSG. As a first phase of load reduction, the controls will begin tomodulate valves (960) and (967) in a closing direction to reduce flowthrough the parallel feedwater heating loop. Once the parallel feedwaterheating loop has been completely isolated, the second phase of controllowers the power output of the GT. Control now returns to the initialprocess block (1602).

[0555] From decision block (1605), if the HRSG required flow is greaterthan the HP economizer optimum flow, then control proceeds to processblock (1620). If the GT is operating at less than full load, the firstphase of control is to increase GT load. Once the GT is operating atfull load, valve (967) is modulated to begin feedwater heating in theparallel loop. Utilizing inputs from the DCS at process block (1610) forthe evaporator section pressure and the temperature exiting the HPeconomizer, valve (967) is modulated to obtain the optimum watertemperature at the exit of the HP economizer. Pump (932) beginsoperation once flow begins to pass through valve (967).

[0556] Control now proceeds to decision block (1621). If the HRSGrequired steam flow is less than the LP economizer optimum flow, thencontrol proceeds to process block (1622). At this power plant load,there is still no need for LP feedwater heating as there is more thansufficient heat available in the exhaust gases to heat the feedwater inthe LP economizer. Therefore, valve (960) is closed. Control returns tothe initial process block (1602).

[0557] From decision block (1621), if the HRSG required steam flow isgreater than the LP economizer optimum flow, then control proceeds toprocess block (1623). At this power plant load, conventional LPfeedwater heating is required as there is insufficient heat available inthe exhaust gases to heat the feedwater in the LP economizer. Therefore,valve (960) is modulated to control flow through the LP economizer toits optimum. Control returns to the initial process block (1602).

[0558] Exemplary Overall Power Plant Control Method

[0559] In providing a control logic for the overall plant, some of themajor objectives include improved efficiency and continuous low emissionlevels. These objectives are best attained by operating the GTs at ornear full load. The control logic for the overall combined cycle controlin this example will focus on these objectives. Obviously, one skilledin the art will recognize that to achieve other objectives, this controlscheme may be easily modified to support other priorities.

[0560] Main Control Loop

[0561] Referencing FIG. 17, the control starts at (1701) and continuesto process block (1702) where the loop control begins. Control thenflows to process block (1704). At this point the controller examinesinputs from process block (1703) which include the current overall plantload and the load reference (desired plant load). Based on these inputs,the controller determines the load change requirements. At decisionblock (1705) the controller examines the need for a change in load. Ifthere is no need to change load, the control is returned to the initialprocess block (1702).

[0562] If a load change is required, control flows to decision block(1706) where it must be determined whether the overall plant load needsto be increased or decreased. If it needs to be increased, processcontrol proceeds to the Increase Power Output subroutine (1708). Anexemplary embodiment of this subroutine is illustrated in the flowchartof FIG. 18. If it needs to be decreased, process control proceeds to theDecrease Power Output subroutine (1707). An exemplary embodiment of thissubroutine is illustrated in the flowchart of FIG. 19.

[0563] Increase Power Output

[0564] Referencing FIG. 18, the Increase Power Output subroutine beginsat step (1801) and proceeds to decision block (1802). If the plant isnot operating in a transition zone of operation (zone where one GT is inthe process of either being brought into or out of service), thenprocess control flows to decision block (1804). Note that in FIG. 43,the transition zone of operation is between 70% and 80% of plant load.This zone range may be varied by one skilled in the art to achieve avariety of plant performance objectives.

[0565] If the plant is operating in a transition zone of operation, thenprocess flows to the Transition Control subroutine, (1805). An exemplaryembodiment of this subroutine is illustrated in the flowchart of FIG.20. Control then returns to the end subroutine process block (1803). Allprocess returns to this block (1803) are returned to subroutine block(FIG. 17, 1708), and finally to the initial process block for overallplant control (FIG. 17, 1702).

[0566] At decision block (1804), if all of the plant's GTs areoperating, then process flow proceeds to decision block (1820). At thisjuncture, the controller determines if all of the GTs are operating atfall load. Since the best method to achieve the objectives is to operatethe GTs at full load, if all GTs are not at full load, control flows toprocess block (1821) where load is increased on one or more GTs. Controlnow returns to the end subroutine process block (1803).

[0567] From decision block (1820), if all GTs are at fall load, thencontrol flows to decision block (1822). This block determines whether ornot either the ST or HRSG is operating at an upper limit. For the HRSG,this is typically the supplemental firing temperature. For the ST, thiswould typically be the inlet pressure. This could also be an operationallimit based upon efficiency or another parameter. If any of these limitsis reached, control flows to process block (1823) which will energize astatus light in the control room indicating to the operators that theplant is at full capacity. Control now returns to the end subroutineprocess block (1803).

[0568] From decision block (1822), if the ST or HRSG is not at an upperlimit, then control flows to process block (1824), where the fuel flowto the HRSGs is increased. Control now returns to the end subroutineprocess block (1803).

[0569] From decision block (1802), if all of the plant's GTs are notoperating, then process flow proceeds to decision block (1810). At thisjuncture, the controller determines if all of the GTs that are currentlyoperating are at full load. Again, since the best method to achieve theobjectives is to operate the GTs at full load, if all GTs are not atfull load, control flows to process block (1811) where load is increasedon one or more GTs. Control now returns to the end subroutine processblock (1803).

[0570] From decision block (1810), if all operating GTs are at fullload, then control flows to decision block (1812). This block determineswhether or not either the ST or HRSG is operating at an upper limit. Inaddition to a temperature or pressure limit, this could also be anoperational limit based upon power plant efficiency or other systemrequirements. If any of these limits are reached, control flows to theTransition Control subroutine, process block (1813). An exemplaryembodiment of this subroutine is illustrated in the flowchart of FIG.20. Control then returns to the end subroutine process block (1803).

[0571] From decision block (1812), if the ST or HRSG is not at an upperlimit, then control flows to process block (1814), where the fuel flowto the HRSGs is increased. Control now returns to the end subroutineprocess block (1803).

[0572] Decrease Power Output

[0573] Referencing FIG. 19, the Decrease Power Output subroutine beginsat (1901) and proceeds to decision block (1902). If the plant is notoperating in a transition zone of operation (zone where one GT is in theprocess of either being brought into or out of service), then processcontrol flows to decision block (1904). Note that in FIG. 43, thetransition zone of operation is between 70 and 80% of plant load.

[0574] If the plant is operating in a transition zone of operation, thenprocess flows to the Transition Control subroutine, (1905). An exemplaryembodiment of this subroutine is illustrated in the flowchart of FIG.20. Control then returns to the end subroutine process block (1903). Allprocess returns to this block (1903) are returned to subroutine block(FIG. 17, 1707), and finally to the initial process block for overallplant control (FIG. 17, 1702).

[0575] At decision block (1904), if all of the plant's GTs areoperating, then process flow proceeds to decision block (1920). At thisjuncture, the controller determines whether or not either the ST or HRSGis operating at a lower limit. For the HRSG and ST, these limits wouldbe determined by the plant engineers who would specify the optimum pointto begin shutdown of a GT. If neither of these limits is reached, thencontrol flows to process block (1921), where the fuel flow to the HRSGsis decreased. Control now returns to the end subroutine process block(1903).

[0576] From decision block (1920), if the GT or HRSG is at a lower limitof operation, then process control proceeds to decision block (1922). Ifthe plant output is greater than the upper limit of the transition zoneof operation, control flows to process block (1924) where load isdecreased on one or more GTs. Control now returns to the end subroutineprocess block (1903).

[0577] From decision block (1922), if the plant output is at the upperlimit of the transition zone of operation, then control flows to (1923),the Transition Control subroutine. An exemplary embodiment of thissubroutine is illustrated in the flowchart of FIG. 20. Control thenreturns to the end subroutine process block (1903).

[0578] From decision block (1904), if all of the plant's GTs are notoperating, then process flow proceeds to decision block (1910). At thisjuncture, the controller determines whether or not either the ST or HRSGis operating at a lower limit. For the HRSG and ST, these limits wouldbe determined by the plant engineers who would specify the optimum pointto begin shutdown of a GT. If neither of these limits is reached, thencontrol flows to process block (1911), where the fuel flow to the HRSGsis decreased. Control now returns to the end subroutine process block(1903).

[0579] From decision block (1910), if the GT or HRSG is at a lower limitof operation, then process control proceeds to decision block (1912). Ifthe plant output is greater than the upper limit of the transition zoneof operation, control flows to process block (1914) where load isdecreased on one or more GTs. Control now returns to the end subroutineprocess block (1903).

[0580] From decision block (1912), if the plant output is at the upperlimit of the transition zone of operation, then control flows to (1913),the Transition Control subroutine. An exemplary embodiment of thissubroutine is illustrated in the flowchart of FIG. 20. Control thenreturns to the end subroutine process block (1903).

[0581] Transition Zone Operation

[0582] Referencing FIG. 20, the Transition Control subroutine begins at(2001) and proceeds to decision block (2002). If a load increase isdesired, the process control proceeds to decision block (2010).

[0583] If the plant is at the lower limit of the transition zone ofoperation, process control proceeds to (2011) and an additional GT isstarted and brought on line. Control then returns to process block(2012). At this point, plant load is modulated by prescribed, programmedoutputs for the GTs and STs for a particular transition zone output.Control now returns to the end subroutine process block (2030).

[0584] If a load decrease is required, the process control proceeds todecision block (2020).

[0585] If the plant is at the lower limit of the transition zone ofoperation, process control proceeds to (2021) and a GT is taken off lineand shutdown. Control now returns to the end subroutine process block(2030).

[0586] If the plant is not at the lower limit of the transition zone ofoperation, process control proceeds to (2022) where load is modulated byprescribed, programmed outputs for the GTs and STs for a particulartransition zone output. Control now returns to the end subroutineprocess block (2030).

[0587] Summary

[0588] The preceding method of controlling the HRSGs and power plant hasillustrated how the teachings of the present invention can beadvantageously applied to power plant operations. It should be notedthat the exemplary system control flowcharts of FIGS. 16-20 may beaugmented or trimmed of steps with no loss in generality or scope ofteachings in regards to the present invention.

[0589] The gist of the present invention is that while a large number ofcontrol schemes may be employed to achieve overall cost andenvironmental savings, the basic use of single (or near single) pressureHRSGs in conjunction with supplemental firing can improve the overalleconomics and environmental costs of existing plant technologies.Furthermore, the novel disclosed method of maximizing power plantoperation over a wide range of load while still maintaining the GTs atfull load operation (as contrasted with the prior art) makes thedisclosed control technique a significant improvement in power plantcontrol system engineering.

PREFERRED SYSTEM CONTEXT OF THE PRESENT INVENTION

[0590] The numerous innovative teachings of the present application willbe described with particular reference to the presently preferredembodiment, wherein these innovative teachings are advantageouslyapplied to the particular problems of a HIGH POWER DENSITY COMBINEDCYCLE POWER PLANT. However, it should be understood that this embodimentis only one example of the many advantageous uses of the innovativeteachings herein. In general, statements made in the specification ofthe present application do not necessarily limit any of the variousclaimed inventions. Moreover, some statements may apply to someinventive features but not to others.

[0591] Retrofit Applications

[0592] Today, many nuclear, coal, and oil-fired power plants are stillin operation. With increasing pressure to be efficient in a competitiveelectrical marketplace, along with environmental concerns for theproduction of greenhouse gases and other pollutants, the retrofit ofthese existing steam turbine power plants to combined cycle power plantsbecomes more and more likely. However, conventional combined cycle powerplants produce steam at three pressure levels, while the existing steamturbines at conventional steam power plants are designed for utilizingonly HP steam.

[0593] In GE informative document GER-3582E (1996), entitled “SteamTurbines for STAG™ Combined Cycle Power Systems”, by M. Boss, the authordescribes a basic difference between a ST in a conventional steam powerplant versus a ST in a conventional combined cycle power application:

[0594] “Mass flow at the exhaust of a combined cycle unit in athree-pressure system can be as much as 30% greater than the throttleflow. This is in direct contrast to most units with fired boilers, whereexhaust flow is about 25% to 30% less than the throttle mass flow,because of extractions from the turbine for multiple stages of feedwaterheating.”

[0595] This stated phenomenon greatly complicates the retrofit ofconventional steam power plants to conventional combined cycle powerplants in the prior art. Since conventional power plants accept steam atthe inlet only, at HP pressure, they are not designed to accept the IPand LP steam produced from conventional combined cycle HRSGs. In orderto be effective, it has already been discussed that conventionalcombined cycle power plants in the prior art have a ST to GT power ratioof approximately 0.5:1. Therefore, to retrofit a 400 MW conventionalsteam power plant to a conventional combined cycle would require 800 MWof GT capacity, bringing the total plant capacity to 1200 MW. Theexisting infrastructure, fuel lines, available real estate, and mostimportantly, high voltage power lines, may not be of sufficient size orrating to allow such an uprate (a 200% increase).

[0596] In addition, to obtain the high levels of efficiency for thecombined cycle from the prior art, the ST would need to be modified toaccept IP and LP steam, and would need to have its entire steam path(internal components including rotating and stationary blades) modified,as the ratio of exhaust steam to throttle steam would change from 0.75in the conventional steam power plant to 1.30 in the conventionalcombined cycle power plant. This is a change of 1.3/0.75 or 1.73. Thisis a major change to the steam path of the ST that is very costly andperhaps even restrictive, as the present turbine casings may not beusable in a redesign. To further complicate matters, much of theexisting equipment at the existing steam power plant (condensers, pumps,piping, etc.) would no longer be correct for the conventional combinedcycle configuration. Items such as feedwater heaters are not even usedin the prior art combined cycle.

[0597] Many of the preferred embodiments of the present invention,however, are an ideal solution to the retrofit option of conventionalsteam power plants to combined cycle technology. Since several of thepreferred embodiments of the present invention specify the production ofprimarily HP steam, this is an ideal option for this retrofit. Thecurrent combined cycle technology produces steam at up to 1800 psig,while a typical utility standard for steam power plants is 2400 psig,one preferred inlet pressure for several of the preferred embodiments.In addition, since the present invention can utilize a higher ST to GToutput ratio (for example, approximately 1.2:1.0), only 330 MW of GTcapacity is required to retrofit a 400 MW conventional steam power plantto become a clean, efficient combined cycle power plant as described byseveral of the preferred embodiments of the present invention. Also,much of the conventional steam power plant equipment, including the ST,feedwater heaters, condenser, pumps, and other auxiliaries could be usedwith little or no modification.

[0598] Retrofit Comparison—Preferred Embodiment to Prior Art

[0599] In U.S. Pat. Nos. 5,375,410 and 5,442,908 Briesch and Costanzorespectively propose a hybrid style power plant suitable for use inretrofit applications, but still utilize a three pressure level HRSG.However, supplemental firing is not utilized, and neither is cooling ofthe HRSG exhaust gases by feedwater. Such retrofit power plants operateas a conventional combined cycle when boiler fuel is not used. Incontrast, the preferred embodiments of the present invention utilizeboiler fuel and/or HRSG supplemental firing to determine the bestbalance between fuel types, fuel economics, part load requirements,and/or plant emission levels.

[0600] An example for comparison of retrofits for existing steam plantsis illustrated in FIG. 44. In this example, an existing steam plantdesigned for standard steam conditions of 2400 psig inlet pressure witha single reheat and inlet/reheat temperatures of 1050° F. is toavailable for retrofit. These steam conditions would normally beassociated with a fossil-fueled power plant, such as coal or oil fired.Although the plant's steam turbine is in good condition, the plant maybe having difficulty with environmental permits, facing expensive boilerrepairs, or be concerned with economic factors in a deregulated powergeneration market. Any one or combination of these factors could beincentive for the plant owners to consider a retrofit of the existingpower plant to the cleaner and more efficient combined cycle technology.

[0601] The conventional steam power plant is rated at 400 MW and has aheat rate of 7620 BTU/kWh. If fuel is expensive, it will be advantageousto upgrade this facility to combined cycle technology. However, thisplant may (partly due to its lower heat rate) have a low appraisedvalue. For this example it is assumed that this plant has a value ofUS$50 million, which equates to only US$125 per kW. With low fuel costs,retrofit may not be economical.

[0602] To design an economical retrofit, it is necessary to select thebest equipment combinations that maximize the ST efficiency andcapability. For a large ST such as the one in this example, itsconstruction would be similar to that shown in FIG. 51. As can be seenfrom this illustration, the rotating and stationary blades in the HP/IPcasing to the left of the figure are much smaller than those in the LPcasings to the right of the figure. Although it is possible to changethe blading in the LP casing, it usually requires a change in the LPcasing, which affects the foundation, support structure, and condensers.The foundation, support structure, and condensers associated with the LPcasings are large heavy components that are difficult and expensive tomodify. Therefore, it is desirable to utilize the ST LP casings withlittle or no modifications, and make steam path changes only to theHP/IP section.

[0603] To maximize the existing ST LP section, it is desirable to matchits exhaust flow in the new combined cycle application to that of theformer steam plant, approximately 1,587,000 lb/hr. Utilizing theindustry standard General Electric (GE) Model PG7241(FA) Gas Turbine asthe GT engine for this uprate, the total steam production from a 3pressure level HRSG used with this GT would only be 528,000 lb/hr.Therefore, 3 GTs of this model would be required in the prior art toeffect this retrofit. This new combined cycle plant from the prior artwould be rated at approximately 800 MW with a heat rate of 6040 BTU/kWh.However, due to the substantially reduced flows in the HP/IP section ofthe existing ST, the blading in these sections would need to bemodified. Also, due to the lower volumetric flows, the ST inlet pressurewould be derated to 1800 psig. The rating of the modified ST would beapproximately 300 MW. Note that since the combined cycle from the priorart doesn't utilize feedwater heaters, these devices would be isolatedfrom service. Total plant modifications, including those to the HP/IPsection of the ST would be extensive and costly, and US$10 million hasbeen allotted to account for these ST modifications.

[0604] Utilizing the technology described by the preferred embodiment onthe present invention, there are at least two options for this retrofit,demonstrating the flexibility that is offered by the invention. Thefirst option utilizes only one industry standard General Electric (GE)Model PG7241(FA) Gas Turbine and HRSG. This option requires a great dealof supplemental firing, but also produces a great deal of steam. Withmatched exhaust flow to the conventional steam plant, the flows to theinlet of the ST are approximately 93% of the conventional steam plantdesign. Therefore, this ST can be used without modification, with only a7% reduction in inlet pressure at rated conditions. In addition, thisdesign will make use of the existing feedwater heaters. The rating ofthe modified ST would be approximately 375 MW, with a total combinedcycle plant heat rate of 6235 BTU/kWh.

[0605] The second option utilizes two industry standard General Electric(GE) Model PG7241(FA) Gas Turbines and HRSGs. With this option, theexhaust flow of the ST exceeds its former design by about 15%.Therefore, the exhaust pressure will climb by about this same amount andoverall efficiency will be decreased. In this option of the preferredembodiment on the present invention, inlet steam flows are 87% of thesteam plant design value, therefore, the ST can be utilized withoutmodification, but with a reduction in inlet pressure at designconditions. This design will also make use of the existing feedwaterheaters. The rating of the modified ST for this second option would beapproximately 395 MW, with a total combined cycle plant heat rate of6060 BTU/kWh.

[0606]FIG. 44 tabulates the data for the various retrofit options. Asdescribed previously, the ultimate determining factor for the retrofitwill be the economic evaluation. If either fuel costs or the plantutilization factor are extremely low, the retrofit may not be warranted.Higher fuel costs may dictate a more efficient plant, but still one withreasonable cost. Limitations on fuel delivery, power line capacity, orreal estate may place restrictions on the power output or the amount ofequipment. Ultimately, the preferred embodiment of the present inventionoffers more options, better utilization of the existing ST, lessinfrastructure change, and lower cost than the retrofit combined cyclepower plant from the prior art.

[0607] Combined Cycle Power Plants

[0608] The present invention is particularly amenable to application incombined cycle power plants, where the current trend is toward gas-firedcombined cycle turbine systems. The features of the present inventionare attractive in this configuration particularly because of the reducedhardware, space, and capital costs using the teachings of the presentinvention. For example, it is entirely feasible using the teachings ofthe present invention to design a high power density combined cyclepower plant having an initial capital cost which is 25% lower than anequivalent prior art combined cycle configuration.

[0609] For example a US$340 million (reference FIG. 25) conventionalcombined cycle power plant from the prior art which can be constructedthrough the methods described by the preferred embodiment of the presentinvention, could be built for US$240 million in (reference FIG. 27)capital costs. Initial savings are US$100 million dollars. These savingsequate to US$10 million annually in financed capital costs assuming an8% interest rate amortized over 20 years. Assuming fuel costs for a 725MW plant from the prior art of US$93.4 million per year (reference FIG.25), the annual savings of US$10 million in capital costs equates to10.7% of the total annual fuel costs for the plant. This means that thepresent invention can be up to 10.7% less fuel efficient than currentcombined cycle configurations and still be more economically viable.Obviously, the goal of the present invention is to be as fuel efficientand as environmentally efficient as possible. Thus, the cost savingsover the life of the plant can be significant.

[0610] In many new power plant constructions or especially in situationswhere the power plant is a retrofit or upgrade to an existinginstallation, the amount of real estate available to construct the newplant is fixed. Thus, the present invention capability of providing anequivalent amount of power output with less plant real estate becomesvery attractive, especially when overall plant efficiencies can bemaintained at or above current levels.

[0611] Furthermore, the ability of the present invention to operateefficiently over a wide range of part loads is a drastic improvementover the prior art, both from a fuel efficiency standard as well as anexhaust emissions standpoint. Finally, the ability of the presentinvention when targeted toward this application to reduce the overallheat rejection of a high capacity power plant is extremely attractive inlight of the negative impact that this waste heat has on theenvironment, especially considering recent scientific studies concerningglobal warming and the like.

[0612] Energy Transport Fluids

[0613] As will be well known by one skilled in the art, while thepreferred embodiments have made use of energy transport fluids (ETF)comprising primarily air in the topping cycle and steam and/or hot waterin the bottoming cycle, the present invention is amenable to applicationwith a wide variety of other energy transport fluids such as ammonia,chlorinated fluorocarbons, oil, etc.

[0614] These are just a few of the exemplary energy transport fluidsthat will work in some context with the present invention, and anymention of “energy transport fluid” or “ETF” should be given itsbroadest meaning when interpreting the intended applications in whichthe teachings of the present invention are germane.

[0615] Combustible Fuel and/or Fuel/Heat Sources p As will be well knownby one skilled in the art, while the preferred embodiments have made useof combustible fuel (CF/CFT/CFB) comprising primarily natural gas, thepresent invention is amenable to application with a wide variety ofother combustible fuels such as hydrocarbon based fuels, fossil fuels,fuel oil, diesel fuel, and jet fuel. Of course, combinations of singlecombustible fuels may be either mixed and fired or fired separately togenerate a hybrid combustible fuel system that would also be within theanticipated scope of the present invention. These are just a few of theexemplary combustible fuels that will work in some context with thepresent invention, and any mention of “combustible fuel” should be givenits broadest meaning when interpreting the intended applications inwhich the teachings of the present invention are germane.

[0616] Similarly, any mention of the term “fuel/heat source (FHS)” whilespecifically including heat generated from the combustion of naturalgas, may also include heat generated from any combustible fuel(CF/CFT/CFB) as defined above, but also may comprise in whole or in partheat derived from a geothermal source, nuclear reactor, nuclear fission,indirect combustion and/or other source of energy.

[0617] GT Engine Availability

[0618] With the onset of electrical deregulation, there has been aflurry of activity by power plant developers to be the first to themarketplace with new capacity. The business strategy for thesedevelopers is that after enough power plants have been constructed in aparticular region, the banks and other financial institutions will bereluctant to finance additional power plants in that region. Therefore,the general consensus seems to be that he who builds his plant first,wins the economic race.

[0619] This rush to the marketplace has had an effect on the GTmanufacturers. At the current time (2nd quarter of 1999) there isapproximately a 3 year wait for a GE frame 7 GT. In recent years, thelead time for one of these GTs was less than 10 months. This is alsonoted in POWER MAGAZINE, (ISSN 0032-5929, March/April 1999, page 13):

[0620] “Gas turbines, which have sold at a modest clip for the past fewyears, suddenly are selling like stocks with a “dot-com” address, asregulated utilities and independent power producers (IPPs) rush todevelop capacity throughout North America. Some companies are placingorders for dozens of turbines, locking up production slots of the majormanufacturers for years to come.”

[0621] This spike in demand for GTs has not only increased the sellingprice of most GTs by a considerable margin, but has made it difficult toeven purchase some models of GTs without a 2-4 year wait for delivery.Therefore, the preferred embodiment of the present invention serves tocircumvent this problem by producing more power in the ST. Thisalleviates the need for such large amounts of GT capacity, and in someexemplary preferred embodiments, twice the capacity can be attainedwhile utilizing the same GTs that would have been used in a combinedcycle from the prior art.

[0622] Westinghouse Model 501G GT Engine

[0623] The Westinghouse model 501G gas turbine engine is the next stepin technology from the “F” class engines (includes GE frame 7FA andWestinghouse 501F). The “G” technology engines have higher pressureratios, more sophisticated turbine blade materials, and a firingtemperature of 2600° F. To avoid serious thermal distortion or otherdamage due to high temperature in the combustor/turbine section of theseGTs, it is necessary to provide steam into the gas turbine for coolingpurposes. Thus, in this new technology, the GT is dependent upon thesteam cycle for proper operation. This equipment arrangement hasprovided for higher overall combined cycle efficiencies at fall loadpower, however, there are numerous drawbacks to this technology. Some ofthese drawbacks are listed below:

[0624] 1. This technology is not yet proven.

[0625] 2. The cycle does not offer a great deal of flexibility, assupplemental firing is limited to less than 10% power augmentation.Additionally, this supplemental firing lowers the overall plantefficiency.

[0626] 3. With the higher combustion temperatures, NOX is more readilyformed, and anticipated NOX levels are 42 PPM on natural gas versus only9 PPM for a GE frame 7FA GT.

[0627] 4. With the integral steam cooling of the model 501 G combustionsection, comes the requirement for ultra pure steam. Since the steamcooling passages in the GT components are small, deposits and build-upthat can result from steam impurities are not tolerable. Therefore,special condensate polishing systems are required to produce this highlypure steam.

[0628] 5. An examination of a heat balance for a model 50IG indicatesthat some of this cooling steam is consumed in the GT (probablytraveling into the turbine section).

[0629] For a 2X1 501G combined cycle plant this appears to be 35,000 to45,000 lb/hr of steam. This increases the make-up water requirements,increases the duty on the condensate polishing systems, and may besubject to increase with time as the small passages which leak thissteam increase in size due to thermal distortion, erosion, or otherfactors, thus degrading the efficiency.

[0630] 6. Most combined cycles operate with a sliding pressure on thesteam cycle to improve efficiency. However, the cooling steam, whichemanates from the IP boiler on the HRSG, must be maintained at nearlyconstant pressure for adequate cooling. This will have detrimentaleffects on efficiency at part load conditions compared to evenconventional combined cycle power plants in the prior art.

[0631] 7. Due to the higher pressure ratio, the model 50IG requires afuel gas pressure of 600 to 650 psig, versus 350 to 370 psig for a GEframe 7FA. Many pipeline companies will not guarantee pressures tosatisfy the model 501G requirements, so fuel gas compressors are neededin many applications where they would not be required for the “F”technology engines.

[0632] 8. These GTs require more than 3 hours to reach full load, versusless than 30 minutes for “F” technology engines. This limits their usein providing peak power demands.

[0633] As can be seen from this list of drawbacks, the newer technologyengines (including the proposed GE “H” technology engines), have a hostof new schemes to enhance combined cycle efficiency by a few percent,but require a vast amount of restrictive, expensive, and complicatedtechnology to achieve these relatively small incremental increases inefficiency. Although the preferred embodiment of the present inventioncan be used with some of these more advanced engines like the model 501G(however, some changes would be required for cooling steam), many of theexemplary preferred embodiments have focused on the GE frame 7FA andother commercial GT systems due to their proven history, simplicity, lowemissions, and improved efficiency when packaged with the cycledescribed by the preferred embodiment of the present invention.

[0634] Combined Cycle Comparison: “G”/“H” GT Technology vs. “F”Technology

[0635] In light of the impending deregulation of the electric powergeneration marketplace and the subsequent competitive economicenvironment that this deregulation will spawn, the electric powergeneration industry has migrated towards a more sophisticated andcomplicated means of power generation. Specifically, “G” and “H” GTtechnologies have become the preferred GT based combined cycles for manyproposed combined cycle power plant installations.

[0636] However, the use of this technology will not be without itsdrawbacks, both economic and environmental. Specifically, the “G” and“H” GT technologies provide less operational flexibility than theirprevious “F” technology counterparts. These newer technologies require amandatory integration of the GT and ST cycles, as the newer GTs requiresteam cooling of internal GT components. Without this ultrapure,precisely metered cooling steam, these GTs will not operate. Therefore,as the combined cycle plant load changes, the steam cycle will not beable to respond as well as even in the prior art, as cooling steamrequirements will dictate the conditions of some steam that is produced.

[0637] For control, these new technologies still focus plant operationon modulation of GT operation to meet plant load requirements, just asin the prior art. However, due to the nature of their integrated cycles,little or no supplemental firing will be allowed using this technology.This characteristic, when coupled with the plant requirements duringpart load conditions, results in substantially decreased part load heatrates even as compared to older “F” technology plants where there is nodirect coupling between the GT cooling and ST operation. Thus, thesenewer technology GTs are generally designed to be base loaded powerplants. This is in contrast to much of the new plant demand load, whichvaries on a daily and seasonal basis.

[0638] Additionally, these newer technology plants have higher firingtemperatures, resulting in the need for more exotic materials in theirconstruction. These higher temperatures therefore lead notably to highermaintenance costs, and also higher NOX emission levels.

[0639] Additionally, these newer GTs to achieve the higher efficiencies,utilize higher engine pressure ratios. This results in the need forhigher natural gas inlet pressures, requiring the addition of fuel gascompressors in many situations. These fuel gas compressors consume agood deal of power, and serve to lower efficiency, increase cost, andreduce reliability of combined cycle power plants.

[0640] In light of the constraints on operational flexibility, part loadefficiency, increased NOX levels, potential fuel gas compressionrequirements, along with the fact that these “G” and “H” technologymachines have not been proven in even short term operation, the presentinvention has focused on the use of older GT technologies such as the“F” technology. In doing so, the present invention permits decoupling ofthe gas turbine and steam turbine cycles while simultaneously allowingthe GTs to operate at peak fuel and emission efficiencies. The presentinvention using “F” technology provides a power plant that drasticallyimproves part load efficiency, improves plant flexibility, lowersemissions, and drastically lowers overall installed plant cost.

[0641] With a heat rate of 6006 BTU/kWh, for a preferred embodiment ofthe present invention, versus 5830 BTU/kWh for Westinghouse “G”technology and 5690 BTU/kWh for GE “H” technology, this represents onlya 3% and a 5.5% increase in efficiency at rated load for these moresophisticated (yet operationally limited) combined cycle plants from theprior art. Given the lower part load efficiencies, added maintenancecosts, increased capital costs, and lack of operational flexibility, itis unlikely that the “G” and “H” technologies (even with theirincrementally higher full load efficiencies) will provide the economicbenefits available via use of the teachings of the present invention asapplied to combined cycle power plants.

[0642] Although the teachings of the preferred embodiments of thepresent invention focus on “F” technology GTs, they may be applied tothe “G” and “H” technologies, but only with careful guidance by the GTmanufacturers. Note, however, that the teachings of the presentinvention do not specifically limit application to any particular GT orGT manufacturer, but are valid throughout the range of commerciallyavailable GTs, as are known by one skilled in the art.

[0643] Preferred Embodiment Plant Design Method

[0644] Since the preferred embodiment consists of a more flexible designfor a combined cycle, it offers high efficiency (both at full and partload), and has significant cost advantages associated its high powerdensity design. This method for selecting the optimum power plant foroperation and financing is described in subsequent sections below.

[0645] Selection

[0646] Referring to the exemplary flowchart of FIG. 47, the processbegins at the start block (4701), and continues to decision block(4702), where it is determined whether to investigate new constructionor the retrofit of an existing plant. If the plant will be newconstruction, process control flows to decision block (4704). If theplant is to be a hybrid fuel design, process control proceeds to theHybrid Fuel Design Subroutine (4705). Otherwise, process controlcontinues to (4706), where the plant developer, using information in(4707) and other information about his proposed power plant site such astransmission line capacity, real estate availability, and the commercialvalue of electricity, will select a desired combined cycle plant rating(CCR).

[0647] Knowing the CCR, the plant developer will proceed to (4708) and,utilizing the input data from (4709), select the GTs for a preferredembodiment combined cycle from a list of selections, such as thatillustrated in FIG. 29 (note that FIG. 29 is only a partial exemplarylist for demonstration purposes). With the GTs selected, the total gasturbine power output, GTP, can be determined. Proceeding to (4710), thesteam turbine power, STP, can be determined as CCR-GTP.

[0648] Knowing GTP and STP, process control flows to (4711) where theSTP/GTP ratio is calculated. Process control now proceeds to (4712)where the desired efficiency and steam conditions are determined basedupon a characteristic curve similar to that illustrated in FIG. 30.Process control now proceeds to (4801) for an economic evaluation of theselected combined cycle.

[0649] Economic Evaluation

[0650] Referring to FIG. 48, the economic evaluation begins at (4801)and proceeds to block (4802), where inputs for the load profile, fueltypes, fuel cost, and other contributing factors listed in (4803) areused to determine fuel costs and average annual specific fuel cost in$/kWh.

[0651] The process continues to (4804), where inputs for the equipmentcost, installation, financing, and other contributing factors listed in(4805) are used to determine capital costs and average annual capitalcost in $/kWh.

[0652] Process flow continues to (4806), where inputs for inventorycost, maintenance, tools, and other contributing factors listed in(4807) are used to determine maintenance costs and the average annualmaintenance cost in $/kWh.

[0653] The process flows to (4808), where inputs for personnel cost,taxes, insurance, and other contributing factors listed in (4809) areused to determine miscellaneous costs and the average annualmiscellaneous costs in $/kWh.

[0654] Utilizing the data for fuel, capital, maintenance, andmiscellaneous costs, along with the factors listed in (4811), a complete“Economic Pro Forma” is determined for the proposed combined cycle plantfrom the preferred embodiment of the present invention.

[0655] The process continues to decision block (4812) to determine ifthe option selected is acceptable. If so, process flows to (4813) wherethis option is compared to other acceptable options. Process controlproceeds to decision block (4814). If the option calculated is proven tobe advantageous over other acceptable options, it becomes the preferredoption and is saved as such in (4815). Process control continues todecision block (4816). If the new option is not preferred, processcontrol continues to decision block (4816), bypassing (4815).

[0656] From decision block (4816), if more options are desired, processcontrol returns to the Design/Financing process (4701) in FIG. 47.Otherwise, process flows to (4817) where the preferred option isselected as the business plan for the combined cycle project and processflow then ends at (4818).

[0657] Retrofit Plants

[0658] Referring to FIG. 49, the Plant Retrofit Process begins at (4901)and proceeds to decision block (4902), where it is determined whetherthe retrofit is for a hybrid fuel plant or not. If the plant is to be ahybrid design, process flows to the Hybrid Fuel Design Subroutine(4903). After return from this subroutine, the process flows to (4904)to determine the plant economics (see FIG. 48).

[0659] If the plant is not a hybrid design, control proceeds to decisionblock (4905). At this juncture, it must be determined if the existing STwill be modified (new steam path) or used “as is”. If is to be modified,the process goes to (4906) where the new ST rating is determinedutilizing inputs from (4907). From here the process returns to (4908).From decision block (4905), if the ST is to be used “as is”, thenprocess control proceeds to (4908).

[0660] Using inputs from (4909), the ST rating in the proposed combinedcycle is determined and the process continues to (4910). With inputs forfuel, capital, and other contributing factors listed in (4911), a ST/GTpower ratio is selected. Proceeding to (4912), utilizing data similar tothat illustrated in FIG. 29, the GTs can be selected. The process nowcontinues to (4801) which is the determination of plant economics (seeFIG. 48).

[0661] Hybrid Fuel Plants

[0662] Hybrid fuel plants can utilize a number of combustible fuels toprovide energy, as well as nuclear, geothermal, or other heat sources.By integrating the combined cycle described by the preferred embodimentof the present invention along with the hybrid fuel cycle, improvedoverall efficiencies and economics are possible.

[0663] Referring to FIG. 50, the Hybrid Fuel Design Procedure begins at(5001). Control flows to decision block (5002) where the process decideswhether the hybrid will use combustible fuel or a heat source likenuclear or geothermal. If combustible fuel is to be used, process flowsto (5005) where the GTs are selected for the hybrid plant based uponrelative cost of fuels, ST size, desired plant rating, and othercontributing factors as indicated in (5006). From here the subroutinereturns to the point of invocation.

[0664] From decision block (5002), if a heat source like nuclear orgeothermal is to be used, the process flows to (5003) where the GTs areselected for the hybrid plant based upon relative cost of fuels, STsize, desired plant rating, and other contributing factors as indicatedin (5004). From here the subroutine returns to the point of invocation.

[0665] Options

[0666] General

[0667] As noted in previous discussion, one of the prime advantages ofthe preferred embodiment of the present invention is flexibility. Thisis not only apparent in the selection of the combined cycle plantrating, but also in its ability to manifest other power solutions suchas the retrofit of existing steam plants or the integration of cycleswith hybrid fuels. Following is a list of other options that can beeffectively utilized in the preferred embodiment of the presentinvention.

[0668] Equipment Arrangement

[0669] In U.S. Pat. No. 5,649,416, James H. Moore describes variousequipment arrangements which include GTs and STs coupled togetherdriving a common generator. Although the arrangements in FIG. 26illustrate the GTs and ST each with its own respective generator, thereis no reason to insist that this arrangement be required. The teachingsof the preferred embodiment are for a new system and method, and theequipment arrangement could very well be as described by Moore in hispatent, or another arrangement if so desired. Thus, any combination ofsingle-shaft system configurations are anticipated by the presentinvention.

[0670] Other Topping/Bottoming Cycles

[0671] The present invention has been discussed primarily with respectto the use of conventional Brayton/Rankine cycles for the combined cycleapplication discussed herein. However, it should be noted that theteachings of the present invention are equally applicable to the use ofother cycles. While there is no practical limit as to what other cyclesmay be utilized within the context of the present invention, it isspecifically anticipated that the GE Kalina cycle (a bottoming cycle)may be particularly amenable to use in conjunction with the presentinvention.

[0672] Thus, for the purposes of this document, the terms “toppingcycle” and “bottoming cycle” should be given their broadest possiblemeanings consistent with the use of Brayton, Rankine, Kalina, and othercycles available to one skilled in the art. Additionally, it should benoted that the present invention specifically anticipates the use ofmultiple cycles within a given combined cycle application.

[0673] Small Steam Turbine Driven BFP

[0674] For illustrative purposes, the boiler feed pumps (BFP) referencedin this disclosure are assumed to be driven by electric motors. However,in larger steam power plants, these pumps are frequently driven by smallsteam turbines, referred to as boiler feed pump turbines (BFPT). TheBPFTs have several advantages over motors, but the primary advantagesare load response and a reduction in exhaust end blade loading.

[0675] Since these BFPTs utilize low pressure steam at their inlets(typically less than 200 psia), they typically consume a fair amount ofsteam. This steam used by the BPFTs equates to a reduction of steam tothe LP section of the main ST. This reduces the loading on the laststage blades and can often increase the efficiency of the cycle.Advanced Steam Conditions

[0676] In U.S. Pat. No. 5,628,183, Rice discusses development work beingconducted in Europe on higher steam temperatures and pressures, and inthe United States through the Department of Energy (DOE) and theElectric Power Research Institute (EPRI). These include work by SolarTurbines on a pilot project designed for higher cycle efficiencies byutilizing 1500° F. ST inlet steam temperatures. Although not proven inlong term reliable service, as these higher steam pressures and/ortemperatures prove reliable, this technology will be easily implementedinto the preferred embodiment of the present invention.

[0677] Advances in GT Technology

[0678] Gas Turbine technology continues to improve with advances likemore efficient compressors, new metallurgy, higher firing temperatures,higher pressure ratios, and other efficiency enhancements. As these GTadvances become available, they should be able to be integrated into thecycle herein described by the preferred embodiment of the presentinvention.

[0679] Non-Corroding LP HRSG Section

[0680] The detrimental effects of GT exhaust gas condensation and itsability to corrode tubes and fins in the HRSG LP section has beendiscussed. One common way to avoid this condensation problem is toprovide preheated feedwater to the HRSG, such that the feedwater issufficiently warm to be above the dew point of the GT exhaust gases andpreclude the formation of moisture on the HRSG heat exchange surfaces.This method has been illustrated in some of the exemplary preferredembodiments, including FIGS. 35 and 39.

[0681] However, another method that can be utilized is the use of anon-corroding material in the HRSG tubes and fins, typically stainlesssteel. This construction method eliminates the need for feedwaterpreheating, and allows for further cooling of the GT exhaust gases, andthus even greater heat recovery of energy from said gases. The drawback,however, is the added cost for the stainless steel material. In manyinstances, this added cost will outweigh the value of the energy saved.But if fuel prices were high, and material costs relatively low, thisoption may be economically viable.

[0682] Combined HP/LP Pump

[0683] In order to produce the required pressures in the steam cycle, apump is typically employed to pump the feedwater to the desiredpressures. In several of the exemplary preferred embodiments, includingFIGS. 9, 15, 35 and 39, dual pumps are indicated for LP and HP service.These pumps may be multiple as illustrated or may be a single pump. Aswith many pumps utilized for this service, they consist of a series ofimpellers that sequentially pressurize the feedwater. A single pumphousing, with extraction ports at the proper “pressure” (impeller)location can provide an intermediate pressure feedwater, while theremainder of the feedwater continues to the HP outlet. Other pumparrangements can also be devised. The intent of the preferred embodimentof the present invention is not to limit the size or style of pump, butto allow the use of any pump or combination thereof that provides therequired service.

[0684] Waste Heat Recovery

[0685] Throughout the discussion of both prior art combined cycle powerplants and the features of the present invention there has been mentionof losses that occur due to equipment inefficiencies in the overallsystem. For example, this might take the form of losses in the generatordue to non-ideal (non-zero) resistance in the generator windings. Ingeneral, most of the system losses in any combined cycle power plant canbe expressed in terms of waste heat, or heat that is generated but notconverted to mechanical or electrical energy. Generator losses, boilerfeed pump losses, lubrication oil losses, ambient GT heat radiationlosses, and ST heat radiation losses are just a few of these waste heatlosses in a conventional combined cycle application. In conventional(prior art) combined cycle arrangements, these waste heat sources aregenerally assumed to be present and not compensated for, as in theseplant configurations the cost of recovering the heat is not economicaland there is little incentive to use this low energy waste heat in auseful application.

[0686] Because of the excess low level heat contained in the GT exhaustgases, the prior art utilized a multi-pressure level HRSG to maximizeheat recovery. Through the use of continuous supplemental firing, theenergy level at the high temperature section of the HRSG equals orexceeds the energy content at the lower temperature section, introducingthe need for ST extraction steam fed feedwater heaters, common Rankinecycle devices not utilized in conventional combined cycles from theprior art.

[0687] With this increased need for low level heat in the preferredembodiment of the present invention, other sources of heat may beutilized. Referring to FIG. 21, these include the gas turbine losses,GTL (2102), steam turbine generator losses, STL (2110), and othermiscellaneous losses. Now low temperature heat such as heat from enginelube oil, generator heat losses, and GT compartment cooling air can allbe used to preheat feedwater and displace the extraction steam used inthe lower temperature feedwater heaters. The use of this heat not onlyimproves the plant heat rate, but reduces the heat rejectionrequirements for the plant.

[0688] The present invention is somewhat unique in these circumstancesbecause these waste heat sources can be used in conjunction withfeedwater heaters (as illustrated in FIG. 15) to add heat to water thatis subsequently superheated within the HRSG. This practical utilizationof feedwater heaters was not possible with the prior art, as the HRSGwas used to provide this function in the prior art, and feedwaterheating would provide no advantage in the prior art combined cycleconfigurations. Thus, the judicious use of feedwater heating withsupplemental firing in some embodiments of the present invention nowprovide a method of efficiently recovering what was in the prior artunrecoverable waste heat.

[0689] It should be noted that the ability to recover this waste heat ina practical manner can be a significant improvement in overall combinedcycle efficiency. Consider, for example, the case in which 1-2% of thewaste heat generated by the system is recovered and put to good use inthe overall combined cycle. Remembering that a large 1000 MW combinedcycle power plant will expend approximately US$175 million annually forfuel means that even a 1% increase in overall cycle efficiency willequate to large savings in fuel (US$1.75 million annually). If thisimprovement can be sustained over a 20-year life cycle of the powerplant, the total fuel savings would be US$35 million. Thus, waste heatrecovery using the present invention represents a new potential forimproving the overall economic efficiency of combined cycle power plantsthat was not a practical possibility using the prior art.

[0690] It should not go unnoticed that the recovery of waste heatrepresents a direct improvement in overall thermal conversion efficiencyin the combined cycle power plant, resulting in a direct reduction inwarming of the atmosphere. Given the increasing concerns regarding theeffect of global warming on our environment, an emphasis on waste heatrecovery by power plant designers should be a concern on par with thereduction of NOX emissions and other forms of pollution. Since it isestimated that over 100,000 MW of additional electric power plantcapacity will be put online in the next decade, the concerns regardingthe waste heat generated by these plants will be worthy of inspection bythose interested in preserving environmental resources. Additionally,since portions of the waste heat generated by combined cycle powerplants is expelled into the environment, there are significant concernsregarding the impact of this waste heat on both plant and animalwildlife.

[0691] Geothermal Plant Augmentation

[0692] The present invention may be amenable in many circumstances whereexisting or proposed geothermal power plants which have a low degree ofefficiency are to be augmented with a gas turbine to either (1)supplement the geothermal energy production to meet the desired loaddemand or (2) replace losses or reduction in geothermal energyproduction for existing geothermal power plants. Since the equipmentproduction for a geothermal installation is relatively fixed, the lossof efficiency or energy production in an existing geothermal power plantmay result in the plant being inefficient to operate. In some cases, thereduction in geothermal energy flow may result in a plant shutdown, asthe amount of power being produced may fall below a critical thresholdfor practical plant operation.

[0693] The present invention can be advantageously applied to thesescenarios in much the same way it is applied to the recovery of wasteheat in a conventional combined cycle power plant. The only differencein this situation is that the ‘waste heat’ used in the present inventionis recovered from a geothermal source. The result of the use of thisgeothermal heat in conjunction with an optimally fired GT results in apower plant that can have a stable power output (regardless of thequality or stability of the geothermal energy source). Since the presentinvention relies heavily on supplemental firing of the HRSG, thegeothermal energy source can in this application be used via heatexchangers to supplant this supplemental firing to the HRSG and thusdisplace the fuel and/or heat normally supplied for this purpose. Thus,as the geothermal energy source declines in output and/or efficiency,this only results in a corresponding increase in supplemental firingfrom other fuel and/or heat sources. The power plant rated outputremains constant, and can even be increased using the retrofit optionsdiscussed elsewhere within this document.

[0694] Cogeneration Applications

[0695] As mentioned previously, the present invention is particularlyapplicable to cogeneration and combined heat and power (CHP)applications in which both shaft drive and heat are utilized within asingle environment, such as a commercial or industrial plant. In suchapplications, a certain amount of heat from a combined cycle plant maybe used for space heating, chemical feedstock processing, pulpprocessing, paper drying, cogeneration, and/or other industrialprocesses and the like.

[0696] The present invention specifically anticipates that the broadestapplication of the teachings of the present invention will be applicableto all forms of cogeneration and CHP applications. As such, the aboveexamples are illustrative only of the range of applications of thepresent invention. Those skilled in the art will no doubt be able toapply the present invention teachings to a wide variety of otherapplications with no loss of generality.

Performance Comparisons

[0697] ST/GT Efficiency Tradeoff

[0698] To overcome the part load issues associated with electricalsystem load fluctuation, several preferred embodiments of the presentinvention utilize the steam turbine (ST) as the prime engine. The ST canreduce load easily by closing inlet valves or modulating inlet pressureto the engine (through a change in the rate of supplemental firing).This has an attenuated effect on part load efficiency as compared to thedilution of firing temperature as experienced by the GT. Additionally,the ST can actually be designed for optimum efficiency at a designatedpart load point, where the gas turbine almost always is most efficientat full load.

[0699] An understanding of the differences between gas turbines andsteam turbines defines the advantages that STs have in operationalflexibility. Gas turbines consist of a compressor section thatcompresses inlet air (usually at ambient conditions) to anywhere from 3to 30 times atmospheric pressure. This air must then travel to thecombustion zone where it is heated through the combustion of fuel tobetween 1600° F. and 2600° F. at full load, depending upon the GTdesign. These hot pressurized gases then expand through a turbinesection in the GT to produce the power that not only drives thecompressor, but also drives an electrical generator. Approximately ⅔ ofthe power developed by this turbine section is required to drive the aircompressor, while the remaining ⅓ is available to drive the electricalgenerator. Due to the complexity of design, which includes matching thecompressor, combustion system, and turbine section to work as anintegrated unit, GTs are very structured machines. Manufacturerstypically have a variety of models of GTs. However, they are designedfor a distinct output or rating. To obtain a custom designed GT isneither feasible nor economical.

[0700] Steam turbines, in contrast, have very flexible designs. Theyrely upon the plant boiler feed pumps to provide pressurized water andthe plant boiler to provide the heat to convert that high-pressure waterinto steam. Therefore, the ST can easily accommodate a change in powerrequirement at the design stage by simply being configured to pass moresteam flow. This is easily accomplished by using incrementally largerstationary and rotating blades in the ST. Typically, a steam turbinedesign engineer can choose from a family of blades in the high-pressure(HP) sections of the ST that may increment as little as 0.25 inches.Blades in the low-pressure (LP) sections of the turbine usually havehigher increment values. Through this design process two different STsmay, for example, have ratings that vary from 100 to 300 MW, and stillfit in essentially the same casing (from the exterior, these twoturbines would look identical). The key difference would be the bladingon the interior of ST and its flow passing capability.

[0701] Additionally, by proper selection of the LP blading, it ispossible to “overload,” from an efficiency viewpoint, the last stageblades at full load. Therefore, at full load, these blades are lessefficient than at part load. Then, when the load is reduced, theefficiency of the LP section actually increases. This design ispreferred for plants that spend a large portion of their operating lifeat part load, but need to reach peak load for short durations ofseasonal peak system demand. It is this flexibility, along with lowmaintenance requirements and proven reliability, that make the STdesirable as the prime engine in a combined cycle power plant.

[0702] Several preferred embodiments of the present invention define asystem whereby the exhaust gases enter into an HRSG as in the prior art.However, these exhaust gases typically contain a great deal of oxygen.In fact, the oxygen content of the air is typically reduced from a valueof 21% in ambient air to a range of about 12% to 15% in a typical GTexhaust at full load. This leaves a great deal of oxygen remaining inthe GT exhaust gases to burn additional fuel. If sufficient quantitiesof fuel are burned, all the steam that would have been produced as lowerpressure steam in the prior art, can be upgraded to HP steam with theproper system modifications as described by the preferred embodiment ofthe present invention. In this manner, the HP steam flow is greatlyincreased, and the ST size relative to the gas turbine(s) (ST/GT powerratio) goes from a nominal 0.5 in the prior art to a value typicallygreater than 1.0. Therefore, rather than being primarily a GT cycle witha ST recovery cycle, the present invention is more like a conventionalsteam plant with additional GT power production and the ducting ofexhaust gases from the GT into the steam power plant's boilers topreheat air and increase boiler efficiency. To maximize efficiency inseveral preferred embodiments of the present invention, a completeintegration of the cycles is required, including utilization of wasteheat, feedwater heating, and implementation of controls to optimize heattransfer.

[0703] Comparison of Prior Art to Exemplary Preferred Embodiments

[0704] As detailed, the prior art combined cycle technology evolved fromsmaller cogeneration plants. In the state-of-the-art combined cyclepower plant from the prior art, the GT exhausts to an HRSG that istypically either two or three pressure levels. The steam from each ofthese pressure levels is then directed to the ST at the appropriatepoint corresponding to the pressure level of the HRSG section.Supplemental firing is utilized as a means to obtain higher output, butthis done only intermittently to meet peak load, and is accomplishedonly with a reduction in thermal efficiency. Primary load control forthe prior art combined cycle power plant is still achieved by modulatingload on the GT(s). Single pressure level HRSGs can also be employed witha corresponding reduction in thermal efficiency. Higher pressure inletsto the steam turbine typically are not justified as the low volumetricflows to the ST offset any cycle gains from higher pressure by reducedturbine efficiency in the HP section.

[0705] With the prior art of combined cycle technology, the plant isprimarily a GT based plant that added an HRSG to recover the waste heat.The ST is then designed to make the best of this recovered heat (whichis converted to steam by the HRSG at multi-pressure levels). The typicalST/GT output ratio for these combined cycle plants is in the 40% to 60%range, with a typical number for a GE S207FA plant being approximately0.57. With the need to utilize HP, IP, and LP steam, the ST hasrelatively low flows in the HP section and higher flows in the LPsection. This reduces HP volumetric efficiency and increases therelative size and cost of the exhaust section(s). Feedwater heating isdone in the HRSG and conventional ST extraction steam fed feedwaterheating is not employed. Preheating of the feedwater from the condensermay be utilized, but the purpose for this process is not to enhanceefficiency but to avoid condensation of water vapor in the exhaustgases. Since water is contained as humidity in the inlet air, and isalso formed as a product of combustion of hydrocarbon fuels, thisincreased concentration of water vapor in the exhaust gases lowers thedew point. Cold feedwater direct from the condenser can causecondensation on the economizer tubes and fins. This condensation hasbeen demonstrated to corrode these fins, lessen the heat exchangeeffectiveness, and cause detrimental effects in HRSG performance. Thusthe use of a feedwater preheater may be utilized in some applications.

[0706] In summary, the combined cycle plant from the prior art isprimarily a GT based plant with the steam cycle designed as a compromisebetween the best cycle efficiency and optimum exhaust gas heat recovery.There are options such as supplemental firing to increase plant ratingby a nominal amount (typically less than 25%), but this additional powercomes with a penalty on plant heat rate. Due to the rigidity of designof the GT, there is little flexibility in the rating or design of thecombined cycle plant from the prior art. In essence, the prior art is arigid power plant design based on the GT engine or set of engines, withan HRSG, and a ST rated nominally at 50% of the GT output. The SToperates in a dependent mode and follows the GT load.

[0707] In several preferred embodiments of the present invention, the GTexhausts to a single pressure level HRSG (or primarily single pressurelevel) that is designed for continuous supplemental firing. Thissupplemental firing increases the steam production by a significantamount, and subsequently increases the feedwater flows such thatadditional pressure levels in the HRSG are not required to cool theexhaust gases to optimum temperature (approximately 180° F.). Feedwaterflows that exceed the optimum flow through the HRSG are directed toconventional ST extraction fed feedwater heaters to improve steam cycleefficiency. Due to the flexibility of design, the combined cycledescribed by several of the preferred embodiments of the presentinvention has a ST/GT output ratio that can vary from approximately 0.75to 2.25. Of course, those skilled in the art will recognize that otheroutput ratios are also possible and within the scope of the teachings ofthe present invention. For most load variations on the plant, the GT(s)remain primarily at or near their most efficient load (100%) and thesupplemental firing rate is modulated to change the ST load.

[0708] In summary, several of the preferred embodiments of the presentinvention are in essence a large central steam power plant similar tothose known in the prior art of steam power plants, with the boilerreplaced by the HRSGs which continuously bum fuel, just like a boiler ina conventional steam plant. However, GTs have now been added to thecycle which provide oxygen rich (12-15%) hot gases to the boiler (HRSG),increasing its efficiency and allowing for the combustion of additionalfuel. Feedwater heating is accomplished in both the HRSG low temperaturesections and in conventional extraction steam fed feedwater heaters. TheST is larger with more mass flow through the HP and IP sections and lessthrough the LP section (steam extracted for feedwater heating reducesexhaust end flow), increasing volumetric efficiency and decreasingrelative exhaust size. Several preferred embodiments of the presentinvention become combined cycle power plants that are more flexible,have improved full and part load efficiency, and are less expensive toconstruct, operate, and maintain.

[0709] Major Equipment Maintenance Costs

[0710] Besides fuel and capital costs, another large expense forcombined cycle power plants is the cost for maintenance, and especiallymaintenance on the major pieces of equipment such as the GT and the ST.These maintenance costs vary with the equipment model, its complexity,and degree of service (high temperature or low temperature, steady orcyclic duty, etc.). Typically maintenance costs are examined on amills/kWh basis, where a mill is US$0.001 or 0.1 cents U.S. currency.Following is a list of expected maintenance costs for some major piecesof equipment, along with the annual expected maintenance costs basedupon a normalized 200 MW output at 70% capacity (1,500,000,000 kWh peryear): Maintenance Rate Annual Cost Description (mills/kWh) (US $) 2400psig ST 0.5 750,000 GE Frame 7FA GT 2.5 3,750,000 Westinghouse 501G GT4.5 6,750,000

[0711] As can be seen from the maintenance numbers, it is much moreexpensive from a maintenance perspective to operate a GT than it is aST. In addition, the advanced technology GT (model 501G) with its higherfiring temperature, single crystalline blades, and steam cooledcombustion section, is also projected to be an expensive piece ofequipment from a maintenance perspective.

[0712] With the prior art, the GTs produce approximately 67% of thepower, while the ST produces 33% (ST/GT output ratio of 0.5:1.0). Sincethe GTs are modulated to change load, and the ST follows, this ratio isfairly constant throughout the load range. Therefore, over a year ofoperation, in a combined cycle power plant in the prior art, the GTmaintenance factor would be applied to 67% of the kWhs produced, and theST maintenance factor to 33% of the kWhs produced.

[0713] With several of the preferred embodiments of the presentinvention, the ratios are not so simple, for as the total plant loadchanges, the ST output is modulated to the greatest extent possible,while the GTs are maintained at or near full load.

[0714] Part Load Efficiency Comparisons

[0715] The teachings of the present invention can be best explained incomparing the performance comparisons illustrated graphically in FIGS.6, 15, 33, and 22-28.

[0716]FIG. 33 graphically illustrates the part load performancedifference between two state-of-the-art conventional combined cyclepower plants and two preferred embodiments of the present invention.This graph illustrates that the present invention performance issignificantly superior to conventional combined cycle power plants atpart load operation. As can be seen from a comparison of the tabulateddata in FIG. 25 and FIG. 27, over the typical operation profile, theexemplary preferred embodiment of the present invention uses less fuel,costs about US$100 million less to construct, and has NOX emissionswhich are less than ⅓ of those in the Westinghouse combined cycle powerplant. Thus, the present invention embodiment illustrated in FIG. 26provides both significant cost savings and simultaneous savings inenvironmental pollution due to reduced NOX emissions. Thischaracteristic is generally a feature of the present invention teachingsand is in essence the best of both worlds—economic efficiency withsimultaneous reduction in pollution.

[0717] To utilize several of the preferred embodiments of the presentinvention, the HRSG must generally be of a more stout construction tohandle the higher pressures and temperatures than required in the priorart. This can be accomplished in numerous ways. First, the use of awater-wall (vertical tubes filled with feedwater) may be needed to linethe combustion area of the HRSG to protect it from the high combustiontemperatures. As an alternative, the exhaust gases could first be cooledthrough the superheater section (to approximately 800° F.), thenreheated to 1600° F. before continuing though the HRSG. Currently, 1600°F. seems to be the upper temperature limit that manufacturers specifyfor standard HRSG construction. Yet another alternative is the use ofdual grids of duct burners in the HRSG. After the GT exhaust gases areheated to 1600° F., they are allowed to cool through the initialsections of the HRSG, then more fuel (heat) is added through combustionat a point downstream. This adds approximately twice the heat as onegrid burner without exceeding any limiting HRSG temperatures.

CONCLUSION

[0718] The present invention permits a wide variety of applications, butit must be noted in summary that the use of the present invention in thefield of combined cycle power plants is particularly advantageous inlight of current trends in the power generation industry. While theprior art has in general taught away from the use of supplemental firingof HRSGs as a means of increasing overall plant efficiency, the presentinvention has embraced this concept.

[0719] Within the context of an overall improvement in systemefficiency, the present invention promotes the use of supplementalfiring not to generate more steam as in past power generationapplications, but the generation of more high quality steam. By this itis meant that by expending additional fuel in the supplemental firing ofthe HRSG it is possible to generate steam which is more energetic andthus capable of more efficiently generating power when used inconjunction with a suitably designed bottoming cycle engine.

[0720] While the present invention when used in the context of powergeneration will require the construction of HRSGs which can sustainhigher temperatures than are currently the norm, the materials toaccomplish this are readily available and both steam plants and HRSGs inthe prior art have demonstrated at these elevated temperature levels.Furthermore, data in this disclosure indicates that in manycircumstances these HRSGs will be smaller than existing units, meaningthat construction and maintenance costs may be comparable to or evenlower than existing units. In addition, the fact that the HRSGs in thisapplication may in many cases be of the single pressure variety may insome circumstances provide some economy in their design andconstruction.

[0721] It must be stressed that the potential energy density of thepresent invention has significant ramifications in regard to the amountof support hardware required to implement a power plant. To achievereasonable overall efficiency, conventional plants make use of a numberof GTs and STs so that when these units are operated at part load thatthe overall system can be operated with reasonable efficiency. This isprimarily because operating the GT at a part load is in general veryinefficient. The present invention circumvents this environmentallydetrimental effect by endeavoring to operate all GTs at their optimalefficiency (both economic and environmental), thus allowing feweroperating GTs to achieve the same overall efficiency and environmentalimpact while consuming fewer support resources.

[0722] These efficiency enhancements of the present invention arefurther supplemented by a mechanism whereby the capacity of the plantmay be temporarily expanded beyond its normal nominal rating, albeit ata lower efficiency level. This extension of the plant rating to supporthigher loads can be a critical factor in the economics of power plantconstruction, because the environmental and logistical hurdles that mustbe overcome to actually construct new power plants are becoming theparamount economic issues barring new plant construction. As such, thepresent invention permits the useful performance of a given power plantconfiguration to be extended beyond that of a conventional power plant,thus permitting the plant rating to be dynamically adapted to meettemporary overload conditions. This capability can have dramaticeconomic cost and environmental savings in that the present inventionpermits the incremental economic and environmental costs to be reducedin the face of a demand for a temporary increase in plant output.

[0723] Finally, it must be stressed that while past power plant designshave endeavored to optimize their operation based on fuel costs alone,the power plants of the future must incorporate and optimize costs ofcapital, environmental impacts, real estate costs, regulatory costs, andthe ever increasing costs of technology and support machinery. It is theintent of the present invention to address all of these factors inunison and obtain an overall plant design that is a cost effective,power efficient, and environmentally friendly method of generatingpower.

[0724] Although a preferred embodiment of the present invention has beenillustrated in the accompanying Drawings and described in the foregoingDetailed Description, it will be understood that the invention is notlimited to the embodiments disclosed, but is capable of numerousrearrangements, modifications, and substitutions without departing fromthe spirit of the invention as set forth and defined by the followingclaims.

What is claimed is:
 1. A combined cycle power plant process comprisingthe steps of: providing a topping cycle fluid (“TCF”) from a toppingcycle engine (“TCE”) to a predominantly single pressure heat recoverydevice (“HRD”); recovering exhaust heat from the TCF in the HRD;supplemental firing the HRD; providing a bottoming cycle fluid (“BCF”)from the HRD to a bottoming cycle engine (“BCE”) at a single pressurelevel; and wherein supplemental firing of the HRD is substantiallycontinuous such that heat added to said TCF by said supplemental firingincreases the flow of the BCF, resulting in an increase in heat recoveryin the HRD when said increased BCF flows return back through said HRD.2. The process of claim 1 further comprising the step of controlling thesupplemental firing such that an exhaust temperature of the HRD ismaintained in a predetermined temperature range.
 3. The process of claim1 further comprising the steps of: diverting a portion of the BCF afterit has exited the BCE to a parallel BCF loop; extracting a portion ofthe BCF from the BCE; and preheating the diverted BCF in the parallelBCF loop using at least one BCF heater that receives heat by BCEextraction.
 4. A combined cycle power plant process comprising the stepsof: operating at least one gas turbine (“GT”), wherein said GT producesexhaust gas; passing said exhaust gas through at least one heat recoverysteam generator (“HRSG”) associated with said GT; supplemental firingsaid HRSG; producing steam at said HRSG using heat from said exhaust gasand said supplemental heat; passing said steam to at least one steamturbine (“ST”); operating said ST; converting said steam to feedwater;passing said feedwater to said HRSG; maintaining the temperature of saidexhaust gas from said HRSG at a predetermined level by controlling therate of supplemental firing of the HRSG when its associated GT load isbetween 0 and about 100% of its nominal power output capacity; andincreasing plant capacity through additional supplemental firing andmaintaining the temperature of the HRSG exhaust gas at saidpredetermined level by diverting feedwater from the HRSG when itsassociated GT load is at about 100% of its nominal power outputcapacity.
 5. A combined cycle power plant process comprising: providinga topping cycle fluid (“TCF”) from a topping cycle engine (“TCE”) to aheat recovery device (“HRD”); recovering exhaust heat from the TCF inthe HRD; substantially continuously supplemental firing the HRD;providing a bottoming cycle fluid (“BCF”) from the HRD to a bottomingcycle engine (“BCE”); and modulating the rate of BCF flow through atleast a portion of the HRD such that an exhaust temperature of the HRDis maintained at a predetermined temperature range, thereby controllingheat recovery from the TCF in the HRD.
 6. The combined cycle power plantprocess of claim 5 wherein said TCE includes at least one gas turbine(“GT”), wherein said BCE includes at least one steam turbine (“ST”),wherein said HRD includes at least one heat recovery steam generator(“HRSG”), and wherein said BCF is steam in said ST and feedwater in atleast a portion of said HRSG.
 7. The combined cycle power plant processof claim 6 wherein modulating the rate of BCF flow includes modulating arate of feedwater flow through said HRSG by controlling thesubstantially continuous supplemental firing of the HRSG, wherein heatadded by supplemental firing increases steam flow from said HRSG to saidST, thereby increasing the rate of feedwater flow through said HRSG. 8.The combined cycle power plant process of claim 6 wherein modulating therate of BCF flow includes modulating a rate of feedwater flow through atleast a portion of said HRSG by diverting feedwater from said HRSG to aparallel feedwater loop and preheating said parallel feedwater using STextraction steam.
 9. The combined cycle power plant process of claim 6wherein said HRSG produces steam at a single pressure.
 10. The combinedcycle power plant process of claim 6 wherein said HRSG produces highpressure steam predominantly at supercritical pressure at rated combinedcycle plant output.
 11. A combined cycle power plant process comprising:providing a topping cycle fluid (“TCF”) from a topping cycle engine(“TCE”) to a predominantly single pressure level heat recovery device(“HRD”); recovering exhaust heat from the TCF in the HRD; supplementalfiring the HRD; providing a bottoming cycle fluid (“BCF”) from the HRDat a predominantly single pressure level to a bottoming cycle engine(“BCE”); and maintaining a TCF exhaust temperature of said HRD in apredetermined temperature range such that heat recovery in said HRD iscontrolled.
 12. The combined cycle power plant process of claim 11wherein said exhaust temperature of said HRD is maintained in saidpredetermined temperature range by controlling BCF flow through at leasta portion of the HRD.
 13. The combined cycle power plant process ofclaim 11 wherein said exhaust temperature of said HRD is maintained insaid predetermined temperature range by controlling supplemental firingof said HRD.
 14. The combined cycle power plant process of claim 13wherein the supplemental firing is substantially continuous.
 15. Thecombined cycle power plant process of claim 13 wherein totalsupplemental firing input at combined cycle plant rated capacity addedis at least about 30% of the energy input to said TCE.
 16. A method ofoperating a combined cycle power plant comprising at least one gasturbine (“GT”), at least one steam turbine (“ST”), and at least one heatrecovery steam generator (“HRSG”), said method comprising: operatingsaid GT to produce shaft work and exhaust gas; passing said exhaust gasthrough said HRSG; adding supplemental heat to said HRSG; producingsteam at said HRSG using heat from said exhaust gas and saidsupplemental heat; passing said steam to said ST; operating said ST toproduce shaft work; converting said steam to feedwater; passing saidfeedwater to said HRSG; and controlling flow of said feedwater throughat least a portion of said HRSG to control heat recovery in said HRSG.17. The method of claim 16 wherein said steam is produced at said HRSGat substantially a single pressure level.
 18. The method of claim 16wherein said steam produced at said HRSG is high pressure steampredominantly at supercritical pressure at rated combined cycle plantoutput.
 19. The method of claim 16 wherein said flow of feedwaterthrough said HRSG is controlled by modulating a rate of adding saidsupplemental heat to said HRSG such that adding said supplemental heatto said exhaust gas produces increased steam flow from said HRSG to saidST, thereby increasing the flow of feedwater.
 20. The method of claim 16wherein said flow of feedwater through at least a portion of said HRSGis controlled by diverting at least some of said feedwater from saidHRSG to a parallel feedwater loop and supplying said diverted feedwaterback to said HRSG.
 21. The method of claim 20 further includingpreheating said feedwater in said parallel feedwater loop using STextraction steam.
 22. The method of claim 20 further includingpreheating said feedwater in said parallel feedwater loop using energyfrom at least one of generator losses and other auxiliaries.
 23. Themethod of claim 17 wherein adding supplemental heat to said HRSGincludes substantially continuously supplemental firing said HRSG. 24.The method of claim 23 wherein said flow of feedwater through said HRSGis controlled by modulating a rate of said supplemental firing such thatadding said supplemental heat to said exhaust gas produces increasedsteam flow from said HRSG to said ST, thereby increasing the flow offeedwater.
 25. The method of claim 16 further including monitoring saidexhaust gas of said HRSG, and wherein said flow of feedwater throughsaid HRSG is controlled to maintain an exhaust temperature in apredetermined temperature range.
 26. The method of claim 25 wherein saidexhaust temperature is about 180° F.
 27. The method of claim 16 whereinsaid flow of feedwater through at least a portion of said HRSG iscontrolled by diverting at least some of said feedwater away from saidHRSG.
 28. The method of claim 16 further including modulating powerplant load to meet system demand.
 29. The method of claim 28 whereinpower plant load is modulated by at least one of controlling addition ofsaid supplemental heat to said HRSG and controlling operation of saidGT.
 30. The method of claim 16 wherein said flow of said feedwaterthrough at least a portion of said HRSG is controlled by modulating therate of adding said supplemental heat when said at least one GTassociated with said HRSG is at or below about 100% of its nominal GTpower output capacity, and wherein said flow of said feedwater throughsaid HRSG is controlled by diverting at least some of said feedwateraway from said HRSG when said at least one GT associated with said HRSGis above about 100% of its nominal GT power output capacity.
 31. Themethod of claim 16 wherein the operational ratio of the rated output ofsaid at least one ST to said at least one GT is greater than about 0.75.32. The method of claim 17 wherein heat recovery in said HRSG iscontrolled such that a ratio of mass flow through said HRSG of saidfeedwater to mass flow through said HRSG of said exhaust gas issubstantially equal to a ratio of heat capacity of said exhaust gas toheat capacity of said feedwater.
 33. The method of claim 16 wherein saidpower plant includes two GTs, two HRSGs, and one ST.
 34. The method ofclaim 16 wherein controlling flow of said feedwater through at least aportion of said HRSG includes modulating said flow of said feedwaterthrough each feedwater section in said HRSG.
 35. A method of operating acombined cycle power plant comprising at least one gas turbine (“GT”),at least one steam turbine (“ST”), and at least one single pressurelevel heat recovery steam generator (“HRSG”), said method comprising:operating said GT to produce shaft work and exhaust gas; passing saidexhaust gas through said HRSG; adding supplemental heat to said HRSG;producing steam at said HRSG using heat from said exhaust gas and saidsupplemental heat, wherein said steam is produced at substantially asingle pressure level; passing said steam to said ST; operating said STto produce shaft work; converting said steam to feedwater; diverting atleast some of said feedwater from said HRSG to a parallel feedwaterloop; preheating said feedwater in said parallel feedwater loop; andsupplying said diverted feedwater back to said HRSG
 36. The method ofclaim 35 wherein adding supplemental heat to said HRSG includessubstantially continuously supplemental firing said HRSG.
 37. The methodof claim 36 wherein a ratio of mass flow through said HRSG of saidfeedwater to mass flow through said HRSG of said exhaust gas issubstantially equal to a ratio of heat capacity of said exhaust gas toheat capacity of said feedwater.
 38. A method of operating a combinedcycle power plant comprising at least one gas turbine (“GT”), at leastone steam turbine (“ST”), and at least one heat recovery steam generator(“HRSG”), said method comprising: operating said GT to produce shaftwork and exhaust gas; passing said exhaust gas through said HRSG;substantially continuously supplemental firing said HRSG(s), whereintotal supplemental firing energy input is at least about 30% of theenergy input to all of said GT(s); producing steam at said HRSG usingheat from said exhaust gas and said supplemental heat; passing saidsteam to said ST; operating said ST to produce shaft work; convertingsaid steam to feedwater; and passing said feedwater to said HRSG. 39.The method of claim 38 wherein said HRSG is a single pressure levelHRSG, and wherein said steam is produced in said single pressure levelHRSG at substantially a single pressure.
 40. A method of operating acombined cycle power plant comprising at least one gas turbine (“GT”),at least one steam turbine (“ST”), and at least one heat recovery steamgenerator (“HRSG”) associated with at least one GT, said methodcomprising: operating said GT to produce shaft work and exhaust gas;passing said exhaust gas through said HRSG; supplemental firing saidHRSG, wherein total supplemental firing energy input is at least about30% of the energy input to its associated GT; producing steam at saidHRSG using heat from said exhaust gas and said supplemental heat;passing said steam to said ST; operating said ST to produce shaft work;converting said steam to feedwater; passing said feedwater to said HRSG;and maintaining an exhaust temperature of said HRSG in a predeterminedtemperature range such that heat recovery in said HRSG is controlled.41. The method of claim 40 further including diverting at least some ofsaid feedwater from said HRSG to a parallel feedwater loop andpreheating said feedwater in said parallel feedwater loop using STextraction steam.